# Research on the Synchronization and Shock Characteristics of an Air Adjustment Mechanism for Cluster-Type DTH Hammers under Partial Loads

Authors: Lei Lou; Meng-ju Chen; Wei-ye Qin; Wan-rong Wu; Hong-yan Rui
Journal: Shock and Vibration
 (2022)

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## Abstract

Cluster down-the-hole (DTH) hammers are often used in large-diameter hard rock drilling operations. Due to complex rock formation conditions, cluster DTH hammers cannot adjust the working airflow in time according to rock formation conditions, resulting in frequent damage to cluster DTH hammers and low work efficiency. Aiming at this problem, this paper proposes an airflow adjustment mechanism as a solution. In this paper, the operation stability and airflow adjustment characteristics of the adjustment mechanism are studied. The impact characteristics and driving force variation characteristics of the adjustment mechanism under different speeds and loading modes are analyzed, and the results show that the movement impact of the adjustment mechanism increases with the increase of the movement speed of the driving actuator under the same loading mode. With the increase of the eccentric load, the critical movement speed of the driving actuator decreases when the adjustment mechanism moves without impact. The test verifies the synchronization characteristics of the adjustment mechanism under different load modes and different control valves. The results show that, under the condition of uniform load, the synchronization accuracy of the dividing/focusing valve and the proportional valve is good. With the increase of the eccentric load difference, the synchronization performance of the dividing/focusing valve becomes worse, and the average error increases. The test results in this paper are of great significance for guiding the engineering design of cluster DTH hammers.


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## Body

## 1. Introduction

High-efficiency drilling technology is widely used in various geological engineering, such as mining, tunneling, and pile foundation construction. The research of high-efficiency drilling and excavation technology mainly includes the research of drilling and excavation equipment, surrounding rock characteristics of excavation holes, excavation process, and geotechnical characteristics. The research on surrounding rock characteristics and geotechnical characteristics of the excavation hole provides a theoretical basis for the development of high-efficiency excavation equipment.The characteristics of the surrounding rock of the excavation hole mainly study the deformation and damage of the surrounding rock, the supporting and anchoring of the surrounding rock, etc. Kien et al. [1] used the 2D FEM RS2 program to analyze the stability of the rock mass around the large underground cavern. The pore water pressure in the foundation pit of the Naples Metro Line 1 tunnel, the displacement of the retaining structure, the ground anchor force, the temperature, and the settlement of the surrounding buildings are measured and studied in detail by Nicotera and Russo [2]. Liu et al. [3] proposed an impact composite rock breaking method and analyzed the influence of impact parameters on the damage characteristics of surrounding rock. Abbasi et al. [4] investigated and studied the diagenetic properties of the rocks in the Ramshir field. Wang et al. [5] analyzed the influence of bolt setting angle, length, and normal load on the shear strength of jointed rock mass. Russo et al. [6] proposed a mixed pile method to solve the reclamation of polluted land and carried out a detailed study of the mechanical behavior of mixed piles. In this paper, the cluster-type down-the-hole (DTH) hammer in the drilling and excavation equipment category is studied, and the above research provides an important scientific basis for this paper.The cluster-type DTH hammer has a large drilling diameter, which requires a large amount of high-pressure gas during impact operations.The flow and pressure of the gas output by the air compressor should be continuously adjusted according to the changes in the different rock formation. When the rock is hard, high-pressure and large-flow gas should be used to achieve improved drilling efficiency. When the rock is soft, low-pressure and small-flow gas can be used to adapt to the changes. At present, when the cluster DTH hammer is used for operation, the pressure and flow of the airflow cannot be adjusted with the rock changes, resulting in frequent damage of the cluster DTH hammer and low operation efficiency. Therefore, this paper proposes an airflow adjustment mechanism as a solution (Figure1).Figure 1
Cluster type DTH hammer model diagram.When a cluster-type DTH hammer is working, too large or too small air flow will cause the cluster DTH hammer failure, resulting in low operating efficiency. How to distribute the airflow reasonably to improve the operation efficiency of the cluster DTH hammer has become extremely important [7]. Aiming at the operational efficiency of a cluster-type DTH hammer, Zhijun et al. [8] analyzed and tested the tunneling technology of large-diameter cluster-type DTH hammers. Chang-Heon Song et al. [9, 10] and others used computational fluid dynamics (CFD) software to study the slag discharge performance of a cluster-type DTH hammer and the flow law of slag particles and conducted a performance test on an optimized DTH hammer. However, the existing cluster type DTH hammers use the drill pipe for air distribution, and the air flow to each DTH hammer cannot be adjusted. The adjustment mechanism designed in this paper is to reasonably adjust the air flow to each DTH hammer for excavation and slag discharge. If the air distribution is uneven, or the adjustment mechanism is unstable, it will seriously affect the operation efficiency of the DTH hammer. Therefore, it is of great significance to study the airflow distribution characteristics and working stability of the airflow adjustment mechanism for the engineering application of the airflow adjustment mechanism.The adjustment mechanism designed in this paper is driven by three oil cylinders. In order to ensure the reasonable distribution of the air flow, the movement of the three oil cylinders needs to be synchronized. The research on multicylinder synchronous mainly focuses on the control valve, control algorithm, and the load characteristics.Wu et al. [11–13] studied the dual-cylinder synchronization control system of the forging hydraulic presses. Yao et al. [14] developed a multisource network hydraulic system with multiple actuators and a corresponding power management strategy to reduce throttling losses and recover energy. The popular intelligent controller was used for hydraulic servo systems by Si et al. [15–17]. Compared with that of PID control, the response speed was increased by 60%, and the tracking error was reduced by 58.43%. Other researchers [18–20] used AMESim software to analyze the influences of the spool diameter, spring stiffness, and orifice diameter of a synchronous valve on the flow characteristics of the valve. Guo et al. [21] study the nonsingular terminal sliding mode control (NTSMC) with active disturbance rejection control (ADRC), and the fourth-order extended state observer was designed to estimate the disturbance of the system in real time. Simulation experiments show that the nonsingular terminal sliding mode-active disturbance rejection control method exhibited better disturbance rejection capacity and a higher tracking accuracy than NTSMC. The above research mainly focuses on the influence of different control strategies and load characteristics on the synchronization characteristics of multiple cylinders under the same control valve. There is less analysis of the motion characteristics of multiple cylinders under different control valves and load characteristics. This paper analyzes the influence of the dividing/focusing valve and the proportional valve on the kinematic characteristics of the regulating mechanism under different load conditions.In order to ensure the normal operation of the cluster DTH hammer, the airflow adjustment mechanism must work stably under different load characteristics, and there is no impact or jamming during operation. The parameters of the mechanism, such as driving speed, driving force, and mechanism clearance, is of great significance to the stable operation of the mechanism. Aiming at the problem regarding the motion impact of the utilized mechanism, Lin et al. [22, 23] proposed a new redundant parallel mechanism and analyzed its speed, acceleration, and driving force. Aiming at the vibration and shock phenomenon caused by the gap in the working process of a ship steering gear transmission mechanism, Huang et al. [24] carried out different control strategies at the moment of starting the transmission mechanism. Hou et al. [25–27] used the Lagrange method to establish a dynamic model of a mechanism with gaps and analyzed the effects of different gap values, driving speeds, and friction factors on the dynamics and impacts of the mechanism. Based on the Newton-Euler method, Chen et al. [28] established a dynamic analysis model for a 3PRS parallel mechanism and studied the influence of friction on the motion process of the parallel mechanism. To improve the smoothness of movement for the traction robot, Bai et al. [29] improved the structure of the traction robot, and a simulation study was carried out on the mechanical properties and dynamic performance with an ADAMS mechanics system. The above research provides a reference for the research of this paper. This paper mainly analyzes the influence of driving speed and load characteristics on the motion stability of the mechanism.In this paper, the effects of different speeds and different loading modes on the impact characteristics and driving force variation characteristics of the adjusting mechanism are analyzed in detail. The dividing/focusing valve and proportional valve are used to control the driving actuator of the adjustment mechanism, and the dynamic characteristics and movement synchronization characteristics of the airflow adjustment mechanism in the uniform load mode and the eccentric load mode are studied. The research results of this paper provide a scientific basis for the engineering application of the airflow adjustment mechanism of the cluster DTH hammer.


## 2. The Principle of Air Flow Adjustment for a Cluster-Type DTH Hammer

The model of a cluster-type DTH hammer is shown in Figure1. It is mainly composed of a DTH hammer, an air flow adjustment mechanism, and a drill rod. The drill rod is a hollow rod that is used to ensure that air flow can pass through the inside of the rod. The air flow adjustment mechanism mainly controls the air flow in the drill rod, thereby realizing the distribution of the air flow.Figure2 is a schematic diagram of the air flow distribution of the cluster-type DTH hammer adjustment mechanism. First, the air compressor is supplied centrally. The drill pipe of the cluster-type DTH hammer has a hollow design, and high-pressure and large-flow gas flows downwards through the drill pipe. When the gas flows through the air flow regulation mechanism, the regulating valve is actuated by the driving oil cylinder, the regulating valve plate opens, and the high-pressure gas flows through the regulating valve to the low-pressure split gas chamber and is discharged through the exhaust hole to carry out the slag washing operation. The pressure and flow rate of the high-pressure air chamber can be adjusted by adjusting the size of the valve opening. The gas discharged through the vent hole can purge the cuttings at the bottom of the hole, and reasonable adjustment of the valve can realize the effective distribution of drilling and rock-carrying airflows.Figure 2
Schematic diagram of the air distribution of the adjustment mechanism.The airflow adjustment mechanism model is shown in Figure3. The inner cylinder of the adjustment mechanism corresponds to the air flow channel inside the drill rod, and the outer cylinder corresponds to the low-pressure diversion cavity of the DTH hammer. When the adjustment mechanism acts under the action of the driving oil cylinder, the adjusting valve is opened, and the high-pressure gas inside the drill pipe is shunted to the low-pressure shunt air chamber, realizing the effective distribution of airflow.Figure 3
Air flow adjustment mechanism diagram.
(a)(b)


## 3. Control System and Mathematical Model of the Adjustment Mechanism

### 3.1. Mathematical Model of Dividing/Focusing Valve Control

Figure4 is the principle diagram of the valve control system of the flow divider and collection valve, which is mainly composed of a motor pump unit, an electromagnetic directional valve, a drive cylinder group, a flow divider collection valve group, a pressure-reducing valve, and a safety valve.Figure 4
The control scheme of the dividing/focusing valve.The structure of the shunting and collection valve is shown in Figure5. An intermediate spring and a hook are located in the middle of the main spool. When the flow is divided, the hook works under the action of the pressure oil, and when the flow is collected, under the action of the pressure oil, the left and right main spools compress the middle spring. The test in this paper adopts the shunt working condition. The right half of Figure 6 is a physical assembly diagram of the shunting and collection valve. The main spool is divided into left and right halves, which are connected by a hook.Figure 5
Dividing/focusing valve assembly diagram.Figure 6
Dividing/focusing valve body and spool.Figure7 is the assembly diagram of the fixed-difference pressure-reducing valve. The fixed-difference pressure-reducing valve is composed of a valve cover, a valve body, a valve sleeve for the pressure-reducing valve, a spring, a valve core, and a lower piston.Figure 7
Assembly drawing of a fixed-differential pressure-reducing valve.The working principle of the fixed-differential pressure-reducing valve is as follows. The oil from portP of the main valve body passes through the oil inlet hole on the valve sleeve of the pressure-reducing valve; one part acts on the upper surface of the lower piston, and the other part flows to the left through orifice a on the valve sleeve. In the side spring cavity, the other part flows to the right spring cavity through orifice b on the valve sleeve. The oil flowing to the left spring cavity flows to the upper surface of the valve core through the orifice, and the oil flowing to the right spring cavity flows down the lower surface of the piston through the orifice. The hydraulic pressure difference between the oil flowing through throttle holes a and b to the left and right spring chambers is only related to the spring force above the valve core, forming a fixed-differential pressure-reducing valve. The opening areas of orifices a and b are consistent with the changes in the system flow rate to ensure that the pressure drop is constant when the flow rate changes suddenly, and the flow rate range that the synchronous valve can adapt to increases.According to Figures5–7 containing the assembly structures of the diverting and collecting valve, the principle of liquid flow control can be drawn, as shown in Figure 8.Figure 8
Schematic diagram of dividing/focusing valve control.The flow equation of orificea is(1)Q1=CdA12P0−P1ρ.The flow equation of orificeb is(2)Q2=CdA22P0−P2ρ.The flow equation of orificec is(3)Q3=CdA32P0−P3ρ.The flow equation of orificee is(4)QA=CdπDx0+x2P1−PAρ.The flow equation of orificek is(5)QB=CdπDx0−x2P1−PBρ.The total flow continuity equation is(6)Q0=Q1+Q2+Q3+V0βedP0ddt.The force equation of the left spool is(7)P1A0−P3A1=md2xdt+Bdxdt+K1−Kx.The force equation of the right spool is(8)P2A0−P3A1=md2xdt+Bdxdt+K1−Kx.In the formulas,Q1, Q2, and Q3 are the flows through orifices a, b, and c, respectively; QA and QB are the outlet flows of the shunting and collecting valve; Q0 is the inlet flow of the shunting and collecting valve; Cd is the flow coefficient of the orifice; A1, A2, and A3 are the throttle areas of orifices a, b and c, respectively; P0 is the inlet pressure of the dividing and collection valve; P1, P2, and P3 are the pressures of the chambers behind orifices a, b, and c, respectively; PA and PB are the outlet pressures of the divider and collector valve; ρ is the oil density; D is the diameter of the valve core; V0 is the inlet volume of the divider and collector valve; βe is the elastic modulus of the oil; x is the movement of the main spool; K is the stiffness of the left and right springs; K1 is the intermediate spring stiffness; and B is the viscous damping coefficient.


### 3.2. Mathematical Model of Proportional Valve Control

Figure9 is a schematic diagram of the proportional valve control system of the regulating valve, which is mainly composed of a motor pump group, a proportional valve group, a driving cylinder group, a proportional pressure-reducing valve, and a safety valve. The driving cylinder group in the figure is composed of 3 driving cylinders, and the proportional valve group composed of 3 proportional valves controls the driving cylinders. The proportional pressure reducing valve can adjust the pressure of the drive cylinder according to the test requirements.Figure 9
The control scheme of the proportional valve.The main spool of the proportional valve has a displacement sensor and a control circuit board, and the opening displacement of the spool can occasionally be detected by the controller. The control block diagram of the 3 driving cylinders of the proportional valve that controls the cluster-type DTH hammer air flow adjustment mechanism is as follows (Figure10):The current control of the proportional valve amplifier is(9)I=KaΔu.The main spool displacement transfer function is(10)xvI=Kas2/ωm2+2ξm/ωms+1.The mathematical transfer function of the valve-controlled drive cylinder is(11)xpxv=KqsAs2/ωh2+2ξh/ωhs+1.The mathematical model of the integral-separation PID controller is as follows:(12)uk=Kpek+Tiβ∑j=0kej+Tdek−ek−1.In the formulas,Ka, I, and Δu are the proportional valve amplification factor, output current, and input voltage, respectively; xp and xv are the displacements of the hydraulic cylinder and the proportional valve core, respectively; and uk is the output signal of the PID controller.Figure 10
Schematic diagram of synchronous control for the proportional valve.


## 3.1. Mathematical Model of Dividing/Focusing Valve Control

Figure4 is the principle diagram of the valve control system of the flow divider and collection valve, which is mainly composed of a motor pump unit, an electromagnetic directional valve, a drive cylinder group, a flow divider collection valve group, a pressure-reducing valve, and a safety valve.Figure 4
The control scheme of the dividing/focusing valve.The structure of the shunting and collection valve is shown in Figure5. An intermediate spring and a hook are located in the middle of the main spool. When the flow is divided, the hook works under the action of the pressure oil, and when the flow is collected, under the action of the pressure oil, the left and right main spools compress the middle spring. The test in this paper adopts the shunt working condition. The right half of Figure 6 is a physical assembly diagram of the shunting and collection valve. The main spool is divided into left and right halves, which are connected by a hook.Figure 5
Dividing/focusing valve assembly diagram.Figure 6
Dividing/focusing valve body and spool.Figure7 is the assembly diagram of the fixed-difference pressure-reducing valve. The fixed-difference pressure-reducing valve is composed of a valve cover, a valve body, a valve sleeve for the pressure-reducing valve, a spring, a valve core, and a lower piston.Figure 7
Assembly drawing of a fixed-differential pressure-reducing valve.The working principle of the fixed-differential pressure-reducing valve is as follows. The oil from portP of the main valve body passes through the oil inlet hole on the valve sleeve of the pressure-reducing valve; one part acts on the upper surface of the lower piston, and the other part flows to the left through orifice a on the valve sleeve. In the side spring cavity, the other part flows to the right spring cavity through orifice b on the valve sleeve. The oil flowing to the left spring cavity flows to the upper surface of the valve core through the orifice, and the oil flowing to the right spring cavity flows down the lower surface of the piston through the orifice. The hydraulic pressure difference between the oil flowing through throttle holes a and b to the left and right spring chambers is only related to the spring force above the valve core, forming a fixed-differential pressure-reducing valve. The opening areas of orifices a and b are consistent with the changes in the system flow rate to ensure that the pressure drop is constant when the flow rate changes suddenly, and the flow rate range that the synchronous valve can adapt to increases.According to Figures5–7 containing the assembly structures of the diverting and collecting valve, the principle of liquid flow control can be drawn, as shown in Figure 8.Figure 8
Schematic diagram of dividing/focusing valve control.The flow equation of orificea is(1)Q1=CdA12P0−P1ρ.The flow equation of orificeb is(2)Q2=CdA22P0−P2ρ.The flow equation of orificec is(3)Q3=CdA32P0−P3ρ.The flow equation of orificee is(4)QA=CdπDx0+x2P1−PAρ.The flow equation of orificek is(5)QB=CdπDx0−x2P1−PBρ.The total flow continuity equation is(6)Q0=Q1+Q2+Q3+V0βedP0ddt.The force equation of the left spool is(7)P1A0−P3A1=md2xdt+Bdxdt+K1−Kx.The force equation of the right spool is(8)P2A0−P3A1=md2xdt+Bdxdt+K1−Kx.In the formulas,Q1, Q2, and Q3 are the flows through orifices a, b, and c, respectively; QA and QB are the outlet flows of the shunting and collecting valve; Q0 is the inlet flow of the shunting and collecting valve; Cd is the flow coefficient of the orifice; A1, A2, and A3 are the throttle areas of orifices a, b and c, respectively; P0 is the inlet pressure of the dividing and collection valve; P1, P2, and P3 are the pressures of the chambers behind orifices a, b, and c, respectively; PA and PB are the outlet pressures of the divider and collector valve; ρ is the oil density; D is the diameter of the valve core; V0 is the inlet volume of the divider and collector valve; βe is the elastic modulus of the oil; x is the movement of the main spool; K is the stiffness of the left and right springs; K1 is the intermediate spring stiffness; and B is the viscous damping coefficient.


## 3.2. Mathematical Model of Proportional Valve Control

Figure9 is a schematic diagram of the proportional valve control system of the regulating valve, which is mainly composed of a motor pump group, a proportional valve group, a driving cylinder group, a proportional pressure-reducing valve, and a safety valve. The driving cylinder group in the figure is composed of 3 driving cylinders, and the proportional valve group composed of 3 proportional valves controls the driving cylinders. The proportional pressure reducing valve can adjust the pressure of the drive cylinder according to the test requirements.Figure 9
The control scheme of the proportional valve.The main spool of the proportional valve has a displacement sensor and a control circuit board, and the opening displacement of the spool can occasionally be detected by the controller. The control block diagram of the 3 driving cylinders of the proportional valve that controls the cluster-type DTH hammer air flow adjustment mechanism is as follows (Figure10):The current control of the proportional valve amplifier is(9)I=KaΔu.The main spool displacement transfer function is(10)xvI=Kas2/ωm2+2ξm/ωms+1.The mathematical transfer function of the valve-controlled drive cylinder is(11)xpxv=KqsAs2/ωh2+2ξh/ωhs+1.The mathematical model of the integral-separation PID controller is as follows:(12)uk=Kpek+Tiβ∑j=0kej+Tdek−ek−1.In the formulas,Ka, I, and Δu are the proportional valve amplification factor, output current, and input voltage, respectively; xp and xv are the displacements of the hydraulic cylinder and the proportional valve core, respectively; and uk is the output signal of the PID controller.Figure 10
Schematic diagram of synchronous control for the proportional valve.


## 4. Stability Test

The stability test is to verify the motion stability of the airflow adjustment mechanism under different loading loads and different motion speeds. Excessive loading and excessive movement speed will cause damage to the adjustment mechanism. Therefore, it is necessary to conduct a detailed study on the movement speed, impact characteristics, and driving force characteristics of the adjusting mechanism under different loads. Taking into account that the actual test may lead to damage the adjustment mechanism, the stability test is carried out with ADAMS software.


### 4.1. Analysis of Speed Characteristics on Adjustment Mechanism

#### 4.1.1. Experimental Principle

The test assembly diagram is shown in Figure11, The adjustment mechanism is mainly composed of drive actuators, rotating mechanism, support rail, and a control ring. The detailed structure of the rotating mechanism is shown in Figure 12. It can be seen from Figure 12 that the driving actuator drives the drive rod to swing up and down, the drive rod drives the shaft to rotate, the shaft drives the rotating rod to swing, and the rotating rod drives the push rod to move up and down. Because the push rod and the control ring are connected by a hinge, the control ring moves up and down under the drive of the push rod.Figure 11
Test assembly diagram.Figure 12
Detailed structure diagram of the rotating mechanism.According to the motion law of the adjustment mechanism, the motion constraints are set in the ADAMS software, and the setting results are shown in Figure13. Drive rod and shaft, shaft and rotating rod are rigidly connected. Rotating rod and push rod, push rod and control ring are connected by ball joint.Figure 13
Constraint setting diagram.Considering the influence of friction factors, it is necessary to define the moving parts according to the contact model. The contact model parameters defined in ADAMS are shown in Table1.Table 1
Contact model parameters.
Parameter nameValueContact stiffness1e5Force index1.5Damping coefficient50Penetration coefficient0.1Static friction coefficient0.3Coefficient of kinetic friction0.2To verify the movement characteristics of the adjustment mechanism at different speeds, the test adopts the speed drive mode, and the Velocity Function Editor in ADAMS is used to realize the speed change of the driver. The speed value is loaded on the drive actuator. The driver continues to accelerate from 0 s to 0.9 s to reach a maximum speed of 48 mm/s and then continues to decelerate to 0 at 1.8 s.Under different eccentric load modes, it is necessary to apply different loads to the loading area of the control ring and monitor its movement characteristics. The loading position and monitoring points of the control ring are shown in Figure14. E, F, and G are detection points. The eccentric load mode settings are shown in Table 2.Figure 14
Load application position and sensor monitoring point on the control ring.Table 2
The eccentric load value.
ModeForce 1 (N)Force 2 (N)Force 3 (N)125002500250022000200030003150020003500


#### 4.1.2. Experimental Results and Discussion

The test results are shown in Figures15–17. Figures 15–17 are the movement speed curves of each monitoring point in the different eccentric load modes. Figure 15 contains the movement speed curve of each monitoring point in the uniform load state. It can be seen from the figure that when the movement speed is below 40 mm/s, the speed curves of monitoring points E, F, and G and the actuator overlap, and no impact fluctuation is observed. When the movement speed exceeds 48 mm/s, the movement speeds of monitoring points E, F, and G have shock fluctuations, and the maximum speed is approximately 64 mm/s.Figure 15
Movement speed of each monitoring point under eccentric load mode 1.Figure 16
Movement speed of each monitoring point under eccentric load mode 2.Figure 17
Movement speed of each monitoring point under eccentric load mode 3.Figure16 shows the movement speed curve of each monitoring point in eccentric load mode 2. It can be seen from the figure that when the movement speed is below 32 mm/s, the speed curves of monitoring points E, F, and G and the actuator overlap, and no impact fluctuation is observed. When the movement speed exceeds 32 mm/s, the movement speeds of monitoring points E, F, and G have shock fluctuations, and the maximum shock speed is approximately 80 mm/s.Figure17 shows the movement speed curve of each monitoring point in eccentric load mode 3. It can be seen from the figure that when the movement speed is below 25 mm/s, the speed curves of monitoring points E, F, and G and the actuator coincide with no fluctuation. When the movement speed exceeds 25 mm/s, the movement speeds of monitoring points E, F, and G fluctuate, and the maximum speed is approximately 100 mm/s.Comparing Figures15–17, the speed impact critical values of different load modes under speed driving conditions can be seen. In the uniform load state (the eccentric load mode 1), the critical value of the velocity impact is 40 mm/s. In eccentric load mode 2, the critical value of the velocity impact is 32 mm/s, and in eccentric load mode 2, the critical value of the velocity impact is 25 mm/s. As the eccentric load increases, the critical impact velocity gradually decreases, and the impact amplitude gradually increases.


### 4.2. Analysis of Impact Characteristics on Adjustment Mechanism

#### 4.2.1. Experimental Principle

An impact force is caused by the discontinuity of the contact, which makes the component contact surface collide. Because the collision time is very short, and the instantaneous impulse is large, a large impact force may be generated, causing damage to the mechanism. The drive system of the adjustment mechanism is shown in Figure18, which is mainly composed of a driving actuator, a driving rod, a rotating shaft, a rotating rod, a pushing rod, bearing 1, bearing 2, and a joint bearing. By analyzing the change of the impact force, the danger of the adjustment mechanism can be predicted in advance. The test is mainly carried out in eccentric load mode 3 (Table 2).Figure 18
Driving system structure diagram.During the experiment, the driving speed of the driving actuator was set to be 35 mm/s and 50 mm, and the constraints of the drive mechanism are set as in Section4.1. Two monitoring points of impact force were set up in the test, namely, test point 1 and test point 2, as shown in Figure 18.


#### 4.2.2. Experimental Results and Discussion

The impact force change characteristics of the bearing 1 (test point 2), the joint bearing between the pushing rod and the control ring at 35 mm/s and 50 mm/s, are studied. The results of the study are shown in Figures19 and 20.Figure 19
Impact force of the bearing 1.Figure 20
Impact force of joint bearing.Figure19 shows the curve of the impact force at the bearing 1. Figure 19 demonstrates that when the driving mechanism moves at 35 mm/s, the force at the bearing 1 is approximately 3500 N, and the bearing force is stable during the whole movement. When the driving mechanism moves at 50 mm/s, the force on the bearing fluctuates, and an impact phenomenon is observed. The maximum impact force is 6000 N, which is approximately 1.7 times the force encountered when the driving mechanism moves at 35 mm/s.Figure20 shows the change curve of the impact force of the joint bearing between the push rod and the control ring. Figure 20 demonstrates that when the driving mechanism moves at a speed of 35 mm/s, the force of the joint bearing is approximately 2450 N. The force is stable. When the driving mechanism moves at 50 mm/s, the force on the joint bearing, and an impact phenomenon is observed. The maximum impact force is 4200 N, which is approximately 1.7 times the force encountered when the driving mechanism moves at 35 mm/s.Comparing Figures19 and 20, it can be seen that, with the increase of the movement speed, the impact force of the main bearing parts of the adjustment mechanism increases. In order to ensure the smooth operation of the adjustment mechanism, the driving speed of the adjustment mechanism should be controlled below 35 mm/s.


### 4.3. Analysis of Driving Force Characteristics on Adjusting Mechanism

#### 4.3.1. Experimental Principle

The test adopts the speed drive mode, and the Velocity Function Editor in ADAMS is used to realize the speed change of the driver. The speed value is loaded on the drive actuator. The speed curve is in the form of a sine curve, the maximum speed is 50 mm/s, and the time for the actuator to reciprocate once is 1.8s; the constraints of the drive mechanism are set as in Section4.1, and the eccentric load mode is shown in Table 2.


#### 4.3.2. Experimental Results and Discussion

The driving force of the driving mechanism is analyzed in the uniform load state (the eccentric load mode 1), eccentric load mode 2 and eccentric load mode 3, and the results are shown in Figures21–23. Figure 21 is a graph showing the change curves of the driving forces on the three driving actuators in the uniform load state. It can be seen from the figure that, in the uniform load state, the driving force curves of the three drive cylinders almost overlap, which indicates that the three drives receive uniform forces in the uniform load state. However, during the movement, the three drivers are subjected to varying degrees of impact due to the nonlinearity of the mechanism and other factors, and the maximum impact force is approximately 3150 N.Figure 21
Driving force curves of driving cylinder under eccentric load mode 1.Figure 22
Driving force curves under eccentric load mode 2.Figure 23
Driving force curves under eccentric load mode 3.Figure22 is a graph showing the change curves of the driving forces on the three driving actuators in the eccentric load mode 2. It can be seen from the figure that, in eccentric load mode 2, due to the eccentric load, the forces of the control ring are uneven, which causes the forces encountered by the driving actuators to be uneven. During the movement of the control ring, the force of actuator 1 is maintained at approximately 2800 N. The force of actuator 2 is approximately 2200 N, and the force of actuator 3 is approximately 2050 N.Figure23 is a graph showing the change curves of the driving forces on the three driving actuators in eccentric load mode 3. It can be seen from the figure that, in eccentric load mode 3, since the control ring moves under a larger load difference, the load differences among the three actuators are larger. The force of actuator 1 is maintained at approximately 3400 N, the force of actuator 2 is approximately 2050 N, and the force of actuator 3 is approximately 1680 N.Comparing Figures21–23, different eccentric load modes have a greater difference of the driving force on the driving actuators than the uniform load. When the load is uniform, the driving forces of the three driving actuators that drive the control ring movement are almost the same. As the eccentric load increases, the differences among the driving forces of the three actuators increase.


## 4.1. Analysis of Speed Characteristics on Adjustment Mechanism

### 4.1.1. Experimental Principle

The test assembly diagram is shown in Figure11, The adjustment mechanism is mainly composed of drive actuators, rotating mechanism, support rail, and a control ring. The detailed structure of the rotating mechanism is shown in Figure 12. It can be seen from Figure 12 that the driving actuator drives the drive rod to swing up and down, the drive rod drives the shaft to rotate, the shaft drives the rotating rod to swing, and the rotating rod drives the push rod to move up and down. Because the push rod and the control ring are connected by a hinge, the control ring moves up and down under the drive of the push rod.Figure 11
Test assembly diagram.Figure 12
Detailed structure diagram of the rotating mechanism.According to the motion law of the adjustment mechanism, the motion constraints are set in the ADAMS software, and the setting results are shown in Figure13. Drive rod and shaft, shaft and rotating rod are rigidly connected. Rotating rod and push rod, push rod and control ring are connected by ball joint.Figure 13
Constraint setting diagram.Considering the influence of friction factors, it is necessary to define the moving parts according to the contact model. The contact model parameters defined in ADAMS are shown in Table1.Table 1
Contact model parameters.
Parameter nameValueContact stiffness1e5Force index1.5Damping coefficient50Penetration coefficient0.1Static friction coefficient0.3Coefficient of kinetic friction0.2To verify the movement characteristics of the adjustment mechanism at different speeds, the test adopts the speed drive mode, and the Velocity Function Editor in ADAMS is used to realize the speed change of the driver. The speed value is loaded on the drive actuator. The driver continues to accelerate from 0 s to 0.9 s to reach a maximum speed of 48 mm/s and then continues to decelerate to 0 at 1.8 s.Under different eccentric load modes, it is necessary to apply different loads to the loading area of the control ring and monitor its movement characteristics. The loading position and monitoring points of the control ring are shown in Figure14. E, F, and G are detection points. The eccentric load mode settings are shown in Table 2.Figure 14
Load application position and sensor monitoring point on the control ring.Table 2
The eccentric load value.
ModeForce 1 (N)Force 2 (N)Force 3 (N)125002500250022000200030003150020003500


### 4.1.2. Experimental Results and Discussion

The test results are shown in Figures15–17. Figures 15–17 are the movement speed curves of each monitoring point in the different eccentric load modes. Figure 15 contains the movement speed curve of each monitoring point in the uniform load state. It can be seen from the figure that when the movement speed is below 40 mm/s, the speed curves of monitoring points E, F, and G and the actuator overlap, and no impact fluctuation is observed. When the movement speed exceeds 48 mm/s, the movement speeds of monitoring points E, F, and G have shock fluctuations, and the maximum speed is approximately 64 mm/s.Figure 15
Movement speed of each monitoring point under eccentric load mode 1.Figure 16
Movement speed of each monitoring point under eccentric load mode 2.Figure 17
Movement speed of each monitoring point under eccentric load mode 3.Figure16 shows the movement speed curve of each monitoring point in eccentric load mode 2. It can be seen from the figure that when the movement speed is below 32 mm/s, the speed curves of monitoring points E, F, and G and the actuator overlap, and no impact fluctuation is observed. When the movement speed exceeds 32 mm/s, the movement speeds of monitoring points E, F, and G have shock fluctuations, and the maximum shock speed is approximately 80 mm/s.Figure17 shows the movement speed curve of each monitoring point in eccentric load mode 3. It can be seen from the figure that when the movement speed is below 25 mm/s, the speed curves of monitoring points E, F, and G and the actuator coincide with no fluctuation. When the movement speed exceeds 25 mm/s, the movement speeds of monitoring points E, F, and G fluctuate, and the maximum speed is approximately 100 mm/s.Comparing Figures15–17, the speed impact critical values of different load modes under speed driving conditions can be seen. In the uniform load state (the eccentric load mode 1), the critical value of the velocity impact is 40 mm/s. In eccentric load mode 2, the critical value of the velocity impact is 32 mm/s, and in eccentric load mode 2, the critical value of the velocity impact is 25 mm/s. As the eccentric load increases, the critical impact velocity gradually decreases, and the impact amplitude gradually increases.


## 4.1.1. Experimental Principle

The test assembly diagram is shown in Figure11, The adjustment mechanism is mainly composed of drive actuators, rotating mechanism, support rail, and a control ring. The detailed structure of the rotating mechanism is shown in Figure 12. It can be seen from Figure 12 that the driving actuator drives the drive rod to swing up and down, the drive rod drives the shaft to rotate, the shaft drives the rotating rod to swing, and the rotating rod drives the push rod to move up and down. Because the push rod and the control ring are connected by a hinge, the control ring moves up and down under the drive of the push rod.Figure 11
Test assembly diagram.Figure 12
Detailed structure diagram of the rotating mechanism.According to the motion law of the adjustment mechanism, the motion constraints are set in the ADAMS software, and the setting results are shown in Figure13. Drive rod and shaft, shaft and rotating rod are rigidly connected. Rotating rod and push rod, push rod and control ring are connected by ball joint.Figure 13
Constraint setting diagram.Considering the influence of friction factors, it is necessary to define the moving parts according to the contact model. The contact model parameters defined in ADAMS are shown in Table1.Table 1
Contact model parameters.
Parameter nameValueContact stiffness1e5Force index1.5Damping coefficient50Penetration coefficient0.1Static friction coefficient0.3Coefficient of kinetic friction0.2To verify the movement characteristics of the adjustment mechanism at different speeds, the test adopts the speed drive mode, and the Velocity Function Editor in ADAMS is used to realize the speed change of the driver. The speed value is loaded on the drive actuator. The driver continues to accelerate from 0 s to 0.9 s to reach a maximum speed of 48 mm/s and then continues to decelerate to 0 at 1.8 s.Under different eccentric load modes, it is necessary to apply different loads to the loading area of the control ring and monitor its movement characteristics. The loading position and monitoring points of the control ring are shown in Figure14. E, F, and G are detection points. The eccentric load mode settings are shown in Table 2.Figure 14
Load application position and sensor monitoring point on the control ring.Table 2
The eccentric load value.
ModeForce 1 (N)Force 2 (N)Force 3 (N)125002500250022000200030003150020003500


## 4.1.2. Experimental Results and Discussion

The test results are shown in Figures15–17. Figures 15–17 are the movement speed curves of each monitoring point in the different eccentric load modes. Figure 15 contains the movement speed curve of each monitoring point in the uniform load state. It can be seen from the figure that when the movement speed is below 40 mm/s, the speed curves of monitoring points E, F, and G and the actuator overlap, and no impact fluctuation is observed. When the movement speed exceeds 48 mm/s, the movement speeds of monitoring points E, F, and G have shock fluctuations, and the maximum speed is approximately 64 mm/s.Figure 15
Movement speed of each monitoring point under eccentric load mode 1.Figure 16
Movement speed of each monitoring point under eccentric load mode 2.Figure 17
Movement speed of each monitoring point under eccentric load mode 3.Figure16 shows the movement speed curve of each monitoring point in eccentric load mode 2. It can be seen from the figure that when the movement speed is below 32 mm/s, the speed curves of monitoring points E, F, and G and the actuator overlap, and no impact fluctuation is observed. When the movement speed exceeds 32 mm/s, the movement speeds of monitoring points E, F, and G have shock fluctuations, and the maximum shock speed is approximately 80 mm/s.Figure17 shows the movement speed curve of each monitoring point in eccentric load mode 3. It can be seen from the figure that when the movement speed is below 25 mm/s, the speed curves of monitoring points E, F, and G and the actuator coincide with no fluctuation. When the movement speed exceeds 25 mm/s, the movement speeds of monitoring points E, F, and G fluctuate, and the maximum speed is approximately 100 mm/s.Comparing Figures15–17, the speed impact critical values of different load modes under speed driving conditions can be seen. In the uniform load state (the eccentric load mode 1), the critical value of the velocity impact is 40 mm/s. In eccentric load mode 2, the critical value of the velocity impact is 32 mm/s, and in eccentric load mode 2, the critical value of the velocity impact is 25 mm/s. As the eccentric load increases, the critical impact velocity gradually decreases, and the impact amplitude gradually increases.


## 4.2. Analysis of Impact Characteristics on Adjustment Mechanism

### 4.2.1. Experimental Principle

An impact force is caused by the discontinuity of the contact, which makes the component contact surface collide. Because the collision time is very short, and the instantaneous impulse is large, a large impact force may be generated, causing damage to the mechanism. The drive system of the adjustment mechanism is shown in Figure18, which is mainly composed of a driving actuator, a driving rod, a rotating shaft, a rotating rod, a pushing rod, bearing 1, bearing 2, and a joint bearing. By analyzing the change of the impact force, the danger of the adjustment mechanism can be predicted in advance. The test is mainly carried out in eccentric load mode 3 (Table 2).Figure 18
Driving system structure diagram.During the experiment, the driving speed of the driving actuator was set to be 35 mm/s and 50 mm, and the constraints of the drive mechanism are set as in Section4.1. Two monitoring points of impact force were set up in the test, namely, test point 1 and test point 2, as shown in Figure 18.


### 4.2.2. Experimental Results and Discussion

The impact force change characteristics of the bearing 1 (test point 2), the joint bearing between the pushing rod and the control ring at 35 mm/s and 50 mm/s, are studied. The results of the study are shown in Figures19 and 20.Figure 19
Impact force of the bearing 1.Figure 20
Impact force of joint bearing.Figure19 shows the curve of the impact force at the bearing 1. Figure 19 demonstrates that when the driving mechanism moves at 35 mm/s, the force at the bearing 1 is approximately 3500 N, and the bearing force is stable during the whole movement. When the driving mechanism moves at 50 mm/s, the force on the bearing fluctuates, and an impact phenomenon is observed. The maximum impact force is 6000 N, which is approximately 1.7 times the force encountered when the driving mechanism moves at 35 mm/s.Figure20 shows the change curve of the impact force of the joint bearing between the push rod and the control ring. Figure 20 demonstrates that when the driving mechanism moves at a speed of 35 mm/s, the force of the joint bearing is approximately 2450 N. The force is stable. When the driving mechanism moves at 50 mm/s, the force on the joint bearing, and an impact phenomenon is observed. The maximum impact force is 4200 N, which is approximately 1.7 times the force encountered when the driving mechanism moves at 35 mm/s.Comparing Figures19 and 20, it can be seen that, with the increase of the movement speed, the impact force of the main bearing parts of the adjustment mechanism increases. In order to ensure the smooth operation of the adjustment mechanism, the driving speed of the adjustment mechanism should be controlled below 35 mm/s.


## 4.2.1. Experimental Principle

An impact force is caused by the discontinuity of the contact, which makes the component contact surface collide. Because the collision time is very short, and the instantaneous impulse is large, a large impact force may be generated, causing damage to the mechanism. The drive system of the adjustment mechanism is shown in Figure18, which is mainly composed of a driving actuator, a driving rod, a rotating shaft, a rotating rod, a pushing rod, bearing 1, bearing 2, and a joint bearing. By analyzing the change of the impact force, the danger of the adjustment mechanism can be predicted in advance. The test is mainly carried out in eccentric load mode 3 (Table 2).Figure 18
Driving system structure diagram.During the experiment, the driving speed of the driving actuator was set to be 35 mm/s and 50 mm, and the constraints of the drive mechanism are set as in Section4.1. Two monitoring points of impact force were set up in the test, namely, test point 1 and test point 2, as shown in Figure 18.


## 4.2.2. Experimental Results and Discussion

The impact force change characteristics of the bearing 1 (test point 2), the joint bearing between the pushing rod and the control ring at 35 mm/s and 50 mm/s, are studied. The results of the study are shown in Figures19 and 20.Figure 19
Impact force of the bearing 1.Figure 20
Impact force of joint bearing.Figure19 shows the curve of the impact force at the bearing 1. Figure 19 demonstrates that when the driving mechanism moves at 35 mm/s, the force at the bearing 1 is approximately 3500 N, and the bearing force is stable during the whole movement. When the driving mechanism moves at 50 mm/s, the force on the bearing fluctuates, and an impact phenomenon is observed. The maximum impact force is 6000 N, which is approximately 1.7 times the force encountered when the driving mechanism moves at 35 mm/s.Figure20 shows the change curve of the impact force of the joint bearing between the push rod and the control ring. Figure 20 demonstrates that when the driving mechanism moves at a speed of 35 mm/s, the force of the joint bearing is approximately 2450 N. The force is stable. When the driving mechanism moves at 50 mm/s, the force on the joint bearing, and an impact phenomenon is observed. The maximum impact force is 4200 N, which is approximately 1.7 times the force encountered when the driving mechanism moves at 35 mm/s.Comparing Figures19 and 20, it can be seen that, with the increase of the movement speed, the impact force of the main bearing parts of the adjustment mechanism increases. In order to ensure the smooth operation of the adjustment mechanism, the driving speed of the adjustment mechanism should be controlled below 35 mm/s.


## 4.3. Analysis of Driving Force Characteristics on Adjusting Mechanism

### 4.3.1. Experimental Principle

The test adopts the speed drive mode, and the Velocity Function Editor in ADAMS is used to realize the speed change of the driver. The speed value is loaded on the drive actuator. The speed curve is in the form of a sine curve, the maximum speed is 50 mm/s, and the time for the actuator to reciprocate once is 1.8s; the constraints of the drive mechanism are set as in Section4.1, and the eccentric load mode is shown in Table 2.


### 4.3.2. Experimental Results and Discussion

The driving force of the driving mechanism is analyzed in the uniform load state (the eccentric load mode 1), eccentric load mode 2 and eccentric load mode 3, and the results are shown in Figures21–23. Figure 21 is a graph showing the change curves of the driving forces on the three driving actuators in the uniform load state. It can be seen from the figure that, in the uniform load state, the driving force curves of the three drive cylinders almost overlap, which indicates that the three drives receive uniform forces in the uniform load state. However, during the movement, the three drivers are subjected to varying degrees of impact due to the nonlinearity of the mechanism and other factors, and the maximum impact force is approximately 3150 N.Figure 21
Driving force curves of driving cylinder under eccentric load mode 1.Figure 22
Driving force curves under eccentric load mode 2.Figure 23
Driving force curves under eccentric load mode 3.Figure22 is a graph showing the change curves of the driving forces on the three driving actuators in the eccentric load mode 2. It can be seen from the figure that, in eccentric load mode 2, due to the eccentric load, the forces of the control ring are uneven, which causes the forces encountered by the driving actuators to be uneven. During the movement of the control ring, the force of actuator 1 is maintained at approximately 2800 N. The force of actuator 2 is approximately 2200 N, and the force of actuator 3 is approximately 2050 N.Figure23 is a graph showing the change curves of the driving forces on the three driving actuators in eccentric load mode 3. It can be seen from the figure that, in eccentric load mode 3, since the control ring moves under a larger load difference, the load differences among the three actuators are larger. The force of actuator 1 is maintained at approximately 3400 N, the force of actuator 2 is approximately 2050 N, and the force of actuator 3 is approximately 1680 N.Comparing Figures21–23, different eccentric load modes have a greater difference of the driving force on the driving actuators than the uniform load. When the load is uniform, the driving forces of the three driving actuators that drive the control ring movement are almost the same. As the eccentric load increases, the differences among the driving forces of the three actuators increase.


## 4.3.1. Experimental Principle

The test adopts the speed drive mode, and the Velocity Function Editor in ADAMS is used to realize the speed change of the driver. The speed value is loaded on the drive actuator. The speed curve is in the form of a sine curve, the maximum speed is 50 mm/s, and the time for the actuator to reciprocate once is 1.8s; the constraints of the drive mechanism are set as in Section4.1, and the eccentric load mode is shown in Table 2.


## 4.3.2. Experimental Results and Discussion

The driving force of the driving mechanism is analyzed in the uniform load state (the eccentric load mode 1), eccentric load mode 2 and eccentric load mode 3, and the results are shown in Figures21–23. Figure 21 is a graph showing the change curves of the driving forces on the three driving actuators in the uniform load state. It can be seen from the figure that, in the uniform load state, the driving force curves of the three drive cylinders almost overlap, which indicates that the three drives receive uniform forces in the uniform load state. However, during the movement, the three drivers are subjected to varying degrees of impact due to the nonlinearity of the mechanism and other factors, and the maximum impact force is approximately 3150 N.Figure 21
Driving force curves of driving cylinder under eccentric load mode 1.Figure 22
Driving force curves under eccentric load mode 2.Figure 23
Driving force curves under eccentric load mode 3.Figure22 is a graph showing the change curves of the driving forces on the three driving actuators in the eccentric load mode 2. It can be seen from the figure that, in eccentric load mode 2, due to the eccentric load, the forces of the control ring are uneven, which causes the forces encountered by the driving actuators to be uneven. During the movement of the control ring, the force of actuator 1 is maintained at approximately 2800 N. The force of actuator 2 is approximately 2200 N, and the force of actuator 3 is approximately 2050 N.Figure23 is a graph showing the change curves of the driving forces on the three driving actuators in eccentric load mode 3. It can be seen from the figure that, in eccentric load mode 3, since the control ring moves under a larger load difference, the load differences among the three actuators are larger. The force of actuator 1 is maintained at approximately 3400 N, the force of actuator 2 is approximately 2050 N, and the force of actuator 3 is approximately 1680 N.Comparing Figures21–23, different eccentric load modes have a greater difference of the driving force on the driving actuators than the uniform load. When the load is uniform, the driving forces of the three driving actuators that drive the control ring movement are almost the same. As the eccentric load increases, the differences among the driving forces of the three actuators increase.


## 5. Synchronization Test

### 5.1. Test Equipment and Loading Principle

The test equipment is shown in Figures24–26, the hydraulic system control principle diagram is shown in Figures 4 and 9, and the load mode of the loading actuator is shown in Table 3. The pressure and flow of the oil output source of the experimental system can be adjusted according to the needs of the test. The hydraulic system is driven by a servo motor, and the speed range of the servo motor is 0–2000r/min. The maximum flow of the hydraulic system is 20 L/min, and the maximum pressure is 10 MPa. The cylinder diameter and rod diameter of the driving actuator are 60 mm and 30 mm, respectively, and the stroke is 120 mm. The control valves, which you can be seen in Figure 25, that drive the actuators are proportional valves and dividing/focusing valves. The model of the proportional valve is DLHZO-T-040. The model of the dividing/focusing valve is 3FJLZ-L8H. The maximum flow of the loading system is 10 L/min, and the maximum pressure is 10 MPa. The cylinder diameter and rod diameter of the loading actuator are 50 mm and 30 mm, respectively.The data acquisition board of the industrial computer is PXI 5630/5632. The industrial computer can control the proportional valve or the dividing/focusing valve to drive the actuator action and, at the same time, detect the displacement of the actuator. Through the collected displacement information, the synchronous motion characteristics of the actuator can be analyzed. The model of the displacement sensor is KTR11-175, and its installation structure is shown in Figure 26. To detect the movement error of the adjustment mechanism, the industrial computer controls the driving actuator to move back and forth in the form of a triangular wave.Figure 24
Experimental system of the adjusting mechanism.Figure 25
Comparison test for the control valve. (a) Proportional valve. (b) Dividing/focusing valve.
(a)(b)Figure 26
The control system of the adjusting mechanism. (a) Control system. (b) Displacement sensor.
(a)(b)Table 3
The eccentric load mode of the loading actuators.
Eccentric load modeActuator 1 (MPa)Actuator 2 (MPa)Actuator 3 (MPa)133323353358Figure27 is a schematic diagram of the loading system for the adjusting mechanism, which mainly consists of a loading actuator group, a solenoid valve group, a pressure-reducing valve group, an oil pump, a motor, and an overflow valve. The loading actuator is controlled by a pressure-reducing valve group and a solenoid valve group. The pressure-reducing valve group can be set to different pressures. The loading actuator mainly sets different resistances for the driving actuator to verify the synchronization characteristics of the driving actuator under different load conditions.Figure 27
Schematic diagram of the loading system on the adjusting mechanism.Figure28 is a layout diagram of the loading actuators. Three loading actuators are arranged at 120° intervals, one end of which is fixed to the top plate, and the other end is fixed to the loading board. The pressures of the pressure-reducing valves of the three loading actuators are set to different loading states. The inner tube is shortened for the convenience of the test, and the control ring (control valve) can be moved up and down under the influence of the drive actuator.Figure 28
Load scheme of the adjusting mechanism.To verify the synchronization characteristics of the drive actuator, the gap between the inner tube of the ajusting mechanism and the control ring is 20 mm. The pressure of the loading cylinder is set by the pressure-reducing valve group in Figure27, and the load modes of the loading actuators are shown in Table 3.The loading actuator is set with three loading modes through the pressure-reducing valve. The first eccentric load mode is the uniform load mode, where the loads of the three loading actuators are all set to 3 MPa. In the second eccentric load mode, the three loading actuators are set to 3 MPa, 3 MPa, and 5 MPa, respectively. In the third eccentric load mode, the three loading actuators are set to 3 MPa, 5 MPa, and 8 MPa, respectively.


### 5.2. Experimental Results and Discussion

The test results are shown in Figures29–31. During the test, to prevent the error caused by the nonlinearity and inertia of the actuators at the start and end positions, the start and end positions of the driving actuators are not analyzed and are only analyzed the reciprocating motion of the actuators from 20 mm stroke to 95 mm stroke.Figure 29
Test Curves of the eccentric load mode 1. (a) Curves of the eccentric load mode 1 for the dividing/focusing valve. (b) Curves of the eccentric load mode 1 for the proportional valve.
(a)(b)Figure 30
Test curves of the eccentric load mode 2. (a) Curves of eccentric load mode 2 for the dividing/focusing valve. (b) Curves of eccentric load mode 2 for the proportional valve.
(a)(b)Figure 31
Test curves of the eccentric load mode 3. (a) Curves of eccentric load mode 3 for the dividing/focusing valve. (b) Curves of eccentric load mode 3 for the proportional valve.
(a)(b)Figure29 contains graphs produced during the equal load test, in which Figure 29(a) shows the curves of the eccentric load mode 1 for the dividing/focusing valve and Figure 29(b) shows the curves of the eccentric load mode 1 for the proportional valve. In the equal load test, the loads of the three loading actuators are all set to 3 MPa. From 20 s to 40 s, the driving actuators are in the extended working condition. From 40 s to 60 s, the drive cylinders are in the retracted working condition. It can be seen from the figure that, within 20 s to 40 s, the three driving actuators simultaneously extend to drive the control ring to move forwards. When the driving actuator reaches 95 mm, the industrial computer controls the proportional valve to reverse, driving actuators move back, and stop at 20 mm. The driving actuators cyclically reciprocate in the form of a triangular wave at 20 to 95 mm. Comparing Figures 29(a) and 29(b), it can be seen that when the three driving actuators cyclically reciprocate in a uniform load state, the synchronization error of the dividing/focusing valve is approximately 2 mm, and the synchronization error of the proportional valve is approximately 1 mm.Figure30 contains the test curves for eccentric load mode 2, in which Figure 30(a) shows the test curves of eccentric load mode 2 for the dividing/focusing valve and Figure 30(b) shows the test curves of eccentric load mode 2 for the proportional valve. Figure 30(a) shows that, at the moment of starting, the displacement curves of the three driving actuators coincide. With the increase in displacement, due to the uneven load, the spool of the dividing/focusing valve automatically adjusts and redistributes the flow under the effect of the eccentric load. Due to the influence of the spool imbalance force, internal leakage, and orifice machining error, the displacement errors of the three driving actuators gradually increase, and when the end point is reached, the displacement errors reach 12 mm.It can be seen from Figure30(b) that when the proportional valve works in eccentric load mode 2, the displacement curves of the actuators have good coincidence, the control accuracies of the proportional valves are high, and the maximum error is only 1.5 mm.Figure31 is the test curves of the eccentric load mode 3, where Figure 3131(a) shows the test curves of the dividing/focusing valve for driving actuators and Figure 31(b) shows the test curves of the proportional valve for driving actuators. Figure 31(a) shows that, within 20 to 40 mm, the displacement errors of the three driving actuators reach 2 mm. Due to the uneven load, as the displacement increases, the displacement errors of the three driving actuators gradually increase. When the end point is reached, the displacement error reaches 16 mm. It can be seen from Figure 31(b) that when the proportional valve works in eccentric load mode 3, the displacement curves of the actuators have good coincidence, the synchronization control accuracy is high, and the maximum error is controlled within 2 mm.Figures32 and 33 contain the average error curves of the dividing/focusing valve and the proportional valve, respectively. Comparing Figures 32 and 33, in the equal load state, the errors of the dividing/focusing valve and the proportional valve are small. The average error of the dividing/focusing valve is approximately 2.5 mm, and the average error of the proportional valve is approximately 1 mm. As the eccentric load value increases, the average error of the dividing/focusing valve gradually increases. When changing from eccentric load mode 2 to eccentric load mode 3, the average error of the dividing/focusing valve increases from 8 mm to 13 mm, and the error increases from 10% to approximately 17%. In eccentric load mode 2 and eccentric load mode 3, the average error of the proportional valve is always controlled within 2 mm, and the synchronization accuracy is high.Figure 32
The average error curves of the dividing/focusing valve.Figure 33
The average error curves of the proportional valve.


## 5.1. Test Equipment and Loading Principle

The test equipment is shown in Figures24–26, the hydraulic system control principle diagram is shown in Figures 4 and 9, and the load mode of the loading actuator is shown in Table 3. The pressure and flow of the oil output source of the experimental system can be adjusted according to the needs of the test. The hydraulic system is driven by a servo motor, and the speed range of the servo motor is 0–2000r/min. The maximum flow of the hydraulic system is 20 L/min, and the maximum pressure is 10 MPa. The cylinder diameter and rod diameter of the driving actuator are 60 mm and 30 mm, respectively, and the stroke is 120 mm. The control valves, which you can be seen in Figure 25, that drive the actuators are proportional valves and dividing/focusing valves. The model of the proportional valve is DLHZO-T-040. The model of the dividing/focusing valve is 3FJLZ-L8H. The maximum flow of the loading system is 10 L/min, and the maximum pressure is 10 MPa. The cylinder diameter and rod diameter of the loading actuator are 50 mm and 30 mm, respectively.The data acquisition board of the industrial computer is PXI 5630/5632. The industrial computer can control the proportional valve or the dividing/focusing valve to drive the actuator action and, at the same time, detect the displacement of the actuator. Through the collected displacement information, the synchronous motion characteristics of the actuator can be analyzed. The model of the displacement sensor is KTR11-175, and its installation structure is shown in Figure 26. To detect the movement error of the adjustment mechanism, the industrial computer controls the driving actuator to move back and forth in the form of a triangular wave.Figure 24
Experimental system of the adjusting mechanism.Figure 25
Comparison test for the control valve. (a) Proportional valve. (b) Dividing/focusing valve.
(a)(b)Figure 26
The control system of the adjusting mechanism. (a) Control system. (b) Displacement sensor.
(a)(b)Table 3
The eccentric load mode of the loading actuators.
Eccentric load modeActuator 1 (MPa)Actuator 2 (MPa)Actuator 3 (MPa)133323353358Figure27 is a schematic diagram of the loading system for the adjusting mechanism, which mainly consists of a loading actuator group, a solenoid valve group, a pressure-reducing valve group, an oil pump, a motor, and an overflow valve. The loading actuator is controlled by a pressure-reducing valve group and a solenoid valve group. The pressure-reducing valve group can be set to different pressures. The loading actuator mainly sets different resistances for the driving actuator to verify the synchronization characteristics of the driving actuator under different load conditions.Figure 27
Schematic diagram of the loading system on the adjusting mechanism.Figure28 is a layout diagram of the loading actuators. Three loading actuators are arranged at 120° intervals, one end of which is fixed to the top plate, and the other end is fixed to the loading board. The pressures of the pressure-reducing valves of the three loading actuators are set to different loading states. The inner tube is shortened for the convenience of the test, and the control ring (control valve) can be moved up and down under the influence of the drive actuator.Figure 28
Load scheme of the adjusting mechanism.To verify the synchronization characteristics of the drive actuator, the gap between the inner tube of the ajusting mechanism and the control ring is 20 mm. The pressure of the loading cylinder is set by the pressure-reducing valve group in Figure27, and the load modes of the loading actuators are shown in Table 3.The loading actuator is set with three loading modes through the pressure-reducing valve. The first eccentric load mode is the uniform load mode, where the loads of the three loading actuators are all set to 3 MPa. In the second eccentric load mode, the three loading actuators are set to 3 MPa, 3 MPa, and 5 MPa, respectively. In the third eccentric load mode, the three loading actuators are set to 3 MPa, 5 MPa, and 8 MPa, respectively.


## 5.2. Experimental Results and Discussion

The test results are shown in Figures29–31. During the test, to prevent the error caused by the nonlinearity and inertia of the actuators at the start and end positions, the start and end positions of the driving actuators are not analyzed and are only analyzed the reciprocating motion of the actuators from 20 mm stroke to 95 mm stroke.Figure 29
Test Curves of the eccentric load mode 1. (a) Curves of the eccentric load mode 1 for the dividing/focusing valve. (b) Curves of the eccentric load mode 1 for the proportional valve.
(a)(b)Figure 30
Test curves of the eccentric load mode 2. (a) Curves of eccentric load mode 2 for the dividing/focusing valve. (b) Curves of eccentric load mode 2 for the proportional valve.
(a)(b)Figure 31
Test curves of the eccentric load mode 3. (a) Curves of eccentric load mode 3 for the dividing/focusing valve. (b) Curves of eccentric load mode 3 for the proportional valve.
(a)(b)Figure29 contains graphs produced during the equal load test, in which Figure 29(a) shows the curves of the eccentric load mode 1 for the dividing/focusing valve and Figure 29(b) shows the curves of the eccentric load mode 1 for the proportional valve. In the equal load test, the loads of the three loading actuators are all set to 3 MPa. From 20 s to 40 s, the driving actuators are in the extended working condition. From 40 s to 60 s, the drive cylinders are in the retracted working condition. It can be seen from the figure that, within 20 s to 40 s, the three driving actuators simultaneously extend to drive the control ring to move forwards. When the driving actuator reaches 95 mm, the industrial computer controls the proportional valve to reverse, driving actuators move back, and stop at 20 mm. The driving actuators cyclically reciprocate in the form of a triangular wave at 20 to 95 mm. Comparing Figures 29(a) and 29(b), it can be seen that when the three driving actuators cyclically reciprocate in a uniform load state, the synchronization error of the dividing/focusing valve is approximately 2 mm, and the synchronization error of the proportional valve is approximately 1 mm.Figure30 contains the test curves for eccentric load mode 2, in which Figure 30(a) shows the test curves of eccentric load mode 2 for the dividing/focusing valve and Figure 30(b) shows the test curves of eccentric load mode 2 for the proportional valve. Figure 30(a) shows that, at the moment of starting, the displacement curves of the three driving actuators coincide. With the increase in displacement, due to the uneven load, the spool of the dividing/focusing valve automatically adjusts and redistributes the flow under the effect of the eccentric load. Due to the influence of the spool imbalance force, internal leakage, and orifice machining error, the displacement errors of the three driving actuators gradually increase, and when the end point is reached, the displacement errors reach 12 mm.It can be seen from Figure30(b) that when the proportional valve works in eccentric load mode 2, the displacement curves of the actuators have good coincidence, the control accuracies of the proportional valves are high, and the maximum error is only 1.5 mm.Figure31 is the test curves of the eccentric load mode 3, where Figure 3131(a) shows the test curves of the dividing/focusing valve for driving actuators and Figure 31(b) shows the test curves of the proportional valve for driving actuators. Figure 31(a) shows that, within 20 to 40 mm, the displacement errors of the three driving actuators reach 2 mm. Due to the uneven load, as the displacement increases, the displacement errors of the three driving actuators gradually increase. When the end point is reached, the displacement error reaches 16 mm. It can be seen from Figure 31(b) that when the proportional valve works in eccentric load mode 3, the displacement curves of the actuators have good coincidence, the synchronization control accuracy is high, and the maximum error is controlled within 2 mm.Figures32 and 33 contain the average error curves of the dividing/focusing valve and the proportional valve, respectively. Comparing Figures 32 and 33, in the equal load state, the errors of the dividing/focusing valve and the proportional valve are small. The average error of the dividing/focusing valve is approximately 2.5 mm, and the average error of the proportional valve is approximately 1 mm. As the eccentric load value increases, the average error of the dividing/focusing valve gradually increases. When changing from eccentric load mode 2 to eccentric load mode 3, the average error of the dividing/focusing valve increases from 8 mm to 13 mm, and the error increases from 10% to approximately 17%. In eccentric load mode 2 and eccentric load mode 3, the average error of the proportional valve is always controlled within 2 mm, and the synchronization accuracy is high.Figure 32
The average error curves of the dividing/focusing valve.Figure 33
The average error curves of the proportional valve.


## 6. Conclusion

The airflow adjustment mechanism of the cluster DTH impactor is designed, and the stability of the adjustment mechanism is studied. The impact characteristics and driving force variation characteristics of the airflow adjustment mechanism under different speeds and loading modes are analyzed. The results show that when the movement speed is greater than 35 mm/s, the speed of the monitoring position fluctuates greatly. As the movement speed increases, the impact force on the main parts of the airflow adjustment mechanism increases. With the increase of the eccentric load difference, the driving force difference of different parts of the adjustment mechanism increases. The maximum eccentric load difference in this paper is 5 MPa. The extreme value of the eccentric load, which caused adjustment mechanism damaged, is needed for further experimental research.The synchronizing characteristics of the airflow regulating mechanism are studied, and the synchronizing characteristics of the airflow regulating mechanism under the control of the dividing/focusing valve and the proportional valve are analyzed. The test results show that, under the condition of equal load, the synchronization accuracy of the dividing/focusing valve and the proportional valve is high. With the increase of the eccentric load, the average error of dividing/focusing valve gradually increases. When changing from eccentric load mode 1 to eccentric load mode 2, the maximum displacement error of the adjustment mechanism controlled by dividing/focusingvalve increases from 10% to 17%, and the average error of the adjustment mechanism controlled by the proportional valve is always within 2 mm. The synchronous control accuracy of proportional valve is higher.It can be seen from the test results that, in order to ensure the reasonable distribution of air flow, the control valve of the adjustment mechanism should choose a proportional valve. In order to ensure the smooth operation of the adjusting mechanism, the speed of the adjusting mechanism should be less than 35 mm/s. When the eccentric load difference is 5 MPa, the adjustment mechanism can operate normally. However, the eccentric load limit difference that causes damage to the adjustment mechanism needs further study. The results of this paper provide a certain scientific basis for the engineering design and application of the airflow adjustment mechanism of the cluster DTH hammer.



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*Source: 9794391-2022-03-31.xml*