Optimization study on valve seating characteristics of hydraulic variable valve system

Abstract

Variable valve technologies, as an advanced category aimed at enhancing internal combustion engine thermal efficiency and emission performance, are highly favored within the automotive industry. This paper proposes a hydraulic variable valve system that can achieve continuous variation in valve phase and lift while maintaining a simple structure. Pre-tests show that at higher engine speeds, the valve seating characteristics are not stable enough. This study believes that in addition to considering the lower seating velocity, the valve seating characteristics also needs to take into account the timeliness of the valve seating and the permissible requirements of hydraulic oil pressure. In order to improve the valve seating performance of the system at higher speed, we first built a simulation model, analyzed the valve seating influencing factors, and then established an optimization model of valve seating characteristics, optimized the parameters of the buffer mechanism and carried out verification tests using a physical platform. The results show that under the premise of meeting the allowable oil pressure, the maximum valve seating velocity is decreased from by 60% from 1.33 m/s to 0.52 m/s, and the seating phase delay is only 3.3°, realizing the smooth seating and effectively increases the applicable speed of the hydraulic variable valve system by 28.5% to above 4500 r/min.

Keywords

hydraulic variable valve system variable valve timing (VVT)internal combustion engine valve seating buffer mechanism optimization

Introduction

In recent years, there has been increasing pressure and challenges for automotive companies. On one hand, emission standards for pollutants from automobiles are becoming stricter. On the other hand, there is a growing expectation for vehicles to have lower fuel consumption while maintaining performance. Consequently, many advanced technical methods have been used to improve internal combustion engine efficiency and combustion performance.^1–8 Research indicates that the implementation of variable valve technology can decrease pumping losses during part-load conditions,^9,10 help achieve internal EGR to reduce NO_x emissions,^11,12 and enhance thermal efficiency and fuel economy.^13,14

A number of variable valve technologies are now widely used. Among them the most used is variable camshaft timing technology, the representative ones of which include Toyota’s VVT-i and Ford’s VCT.^15,16 However, this technology is unable to regulate the duration angle of valve opening and the maximum lift with changes in engine speed. Coupled with the VANOS (or variable Nockenwellensteuerung), the Valvetronic system by BMW accomplishes continuously variable valve timing and maximum lift, aligning well with valve specifications. However, its structure is intricate and comes with a high cost. Honda’s VTEC, Toyota’s VVTL-i, Audi’s AVS system, and Mitsubishi’s MIVEC all adopt variable cam-line technology, usually consisting of two or three cams. The technology can only meet two or three speeds of optimal valve and still cannot meet the requirements for continuous variable valve at different speeds. The Active Valve Train (AVT) developed by Lotus Engineering is an electro-hydraulic system with the ability to regulate valve timing and maximum lift according to engine operating conditions.^17–19 But each valve requires individual control, which is complex, costly, and requires high electromagnetic response speed. Therefore, existing technologies often suffer from reasons such as structural complexity, high cost, and limited variability of the valve, which limits the improvement effect of variable valve technology on the engine.

With a simple structure, the hydraulic variable valve system studied in this paper is able to achieve continuously variable valve phase and lift compared to other existing variable valve technologies, which is of great significance to further improve engine performance.²⁰ Pre-tests show that the valve seating performance is not stable enough at higher engine speeds, as manifested by a noticeable rebound phenomenon with high seating velocity. Adding a buffer mechanism to the hydraulic system is an effective solution to reduce the valve seating velocity. Jin et al.²¹ designed a thin-walled small hole structure to reduce the valve seating speed and believed that increasing the thin-walled hole diameter would help reduce pressure fluctuations in the hydraulic system while meeting the seating speed requirements. Liu et al.²² used a tapered buffer block to reduce the flow area of the valve at the end of closing to achieve the purpose of reducing the valve seating speed. Tu et al. studied the buffering process of the throttle valve, optimized the throttle area and throttle length, and reduced the valve seating speed to 0.07 m/s. Chen et al.^23,24 analyzed the influence of different orifice shapes on the valve seating performance, and adopted the method of orifice throttling to reduce the maximum valve seating speed to 0.5 m/s.

Based on the analysis of the valve seating influencing factors by using the built simulation model, we established an optimization model and carried out verification tests using a physical platform. The study outcomes enabled the system to meet the requirements for the timeliness of the valve seating and allowable oil pressure, while effectively reducing the valve seating velocity. The research results are of great significance in enhancing the adaptability of the hydraulic variable valve system at high engine speeds.

This paper firstly introduces the working principle and test platform of the hydraulic variable valve system (HVVS), then presents the requirements for valve seating characteristics, and develops a mathematical model and a simulation model of valve seating. Finally, an optimization model of the valve seating buffer mechanism is established, and the verification test is carried out, which effectively improves the valve seating performance.

Hydraulic variable valve system (HVVS)

System structure and working principle

Figure 1 illustrates the schematic diagram outlining the structure of HVVS. This system comprises the driving cam, hydraulic cylinder, valve assembly, regulator, seating buffer, oil supply device, and oil circuit. The working principle is as follows: when the cam initiates the lift section, the system oil pressure gradually increases. Given that the design value of the initial oil pressure corresponding to the regulator spring preload is less than that of the valve spring, the regulator plunger is propelled by hydraulic oil until it reaches the predetermined position. Throughout this phase, the valve remains stationary. As the cam continues its rotation, the valve initiates lifting only when the system oil pressure attains the preload pressure corresponding to the valve. This delays the opening of the valve.

Figure 1.

Schematic diagram of the structure of hydraulic variable valve system.

As the cam transitions into the return stroke phase, the regulator plunger holds its position in the limit until the valve reverses its motion and completes its seating. Only then does the regulator plunger commence its reverse movement back to the initial position. This procedural sequence brings about a change in the valve’s closing phase angle. Additionally, as a portion of the oil flows from the cam chamber into the regulator chamber, inducing a change in the volume entering the valve chamber, the corresponding maximum lift of the valve undergoes variation.

Clearly, by adjusting the predetermined position of the regulator plunger, the HVVS can achieve continuous variability in both valve phase and lift.

System testbed

Figure 2 shows a physical test rig dedicated to the HVVS. Engine speed is simulated using a variable-frequency motor, propelling the valve cam through a 1:1 belt drive to facilitate the valve’s opening. The signal collector gathers real-time data from laser displacement sensors, high-frequency pressure sensors, and rotary encoders, monitoring valve displacement, system pressure, and cam phase. This data is then processed and analyzed using a computer.

Figure 2.

Physical test rig for hydraulic variable valve system.

Figure 3 illustrates the valve test curves at an engine speed of 1500 r/min. In this context, A represents the regulator’s adjustment amount. It is evident that changes in the adjustment amount led to variations in the opening phase angle, closing phase angle, and maximum lift of the valve.

Figure 3.

Valve adjustment test curves.

Requirement of valve seating characteristics

Higher valve seating velocity will cause serious wear on the valve head and valve seat ring, making the valve stem prone to fracture, as well as producing greater impact noise. In addition, excessive valve seating velocity will often cause the valve to rebound, resulting in poor valve closure. To ensure the valve performance and service life, the seating velocity in terms of valve motion must not exceed 0.6 m/s ∼ 0.8 m/s.²⁵

Figure 4 shows the partial enlargement of the valve lift test curves at an engine speed of 4500 r/min. We can see that there is a prominent phenomenon of valve seating rebound. The valve seating rebound height corresponding to different adjustment amounts is 0.14 mm to 0.33 mm, and the valve cannot achieve smooth seating.

Figure 4.

Valve movement pattern test curves at an engine speed of 4500 r/min.

For the HVVS, in addition to the above influences, valve seating also needs to factor in the permissible valve cylinder oil pressure and the timeliness of the valve seating requirements. If there is a significant delay in the valve seating, it may cause intake air reflux, affecting the normal operation of the engine. Therefore, to improve the actual performance of the HVVS, it is important to establish a simulation model based on the mathematical model and analyze the valve seating characteristics and influencing factors for selecting the buffer mechanism parameters of valve seating and to ensure smooth valve seating.^26,27

Construction of mathematical model and simulation model for valve seating

Mathematical model for valve seating

Figure 5 depicts the structure of the valve buffer mechanism. Throughout the process of valve opening, hydraulic oil flows into the valve oil chamber through the side oil-way and a one-way valve, driving the valve movement, and the buffer mechanism remains inactive. During the valve closing process, the one-way valve remains in the closed position, and the oil flows back solely through the side oil passage. As the valve continues its motion, the buffer mechanism gradually comes into effect, creating a throttling effect by narrowing the gap. This results in a reduction of the valve seating velocity. To analyze the seating characteristics of the valve, a mathematical model for the valve buffering process will be established in the following steps.

Figure 5.

Structure diagram of the valve buffer mechanism.

In Figure 6, $l_{c}$ denotes the cone length, $x_{lap}$ represents the distance from the base of the cone to the end of the buffer hole, $α$ is the cone angle, $d_{h}$ signifies the inner diameter of the buffer hole, $d_{a}$ represents the relative plunger diameter, $d_{q}$ is the diameter of the valve plunger, and $c_{r} = \frac{d_{h} - d_{q}}{2}$ designates the minimum clearance within the buffer. When considering the boundary, the oil pressure near the buffer hole side is denoted as $p_{q}$ , while on the opposite side, it is represented as $p_{z}$ .

Figure 6.

Partial diagram of the buffer mechanism.

According to the different formulas for calculating the flow cross-sectional area between the valve plunger and the buffer hole $A_{q}$ , the valve buffering process can be divided into three stages—a, b, and c as shown in Figure 7. The dotted lines are the flow cross-sections at different stages. The three endpoints are defined as E, F, and G in the figure. $x_{l a p} = x_{1}$ , $x_{l a p} = x_{2}$ , and $x_{l a p} = 0$ are the interruption points of the three stages of the buffering process, respectively, where $x_{l a p} = x_{1}$ corresponds to the buffering start point and $x_{l a p} = x_{2}$ corresponds to $∠ E F G = 9 0^{\circ}$ . Following is the construction of the flow equation in the valve cylinder and the force balance equation in the valve assembly for each stage.

Figure 7.

Schematic diagram of the three stages of the valve buffering process. (a, b, c) represent the three stages of buffering respectively.

When $x_{2} < x_{l a p} < x_{1}$

x_{1} = l_{c} + \tan α \cdot (c_{r} + l_{c} \tan α)

(1)

x_{2} = l_{c} + \sqrt{[{(\frac{\frac{π}{4} d_{h}^{2}}{π \cdot (d_{q} + c_{r} - l c \cdot \tan α)})}^{2} - {(c_{r} + l c \cdot \tan α)}^{2}]}

(2)

A small orifice throttle exists between the buffer plunger and the buffer orifice, with the flow cross-sectional area of $A_{q}$ being the side area of the circular table where the line EF is located.

A_{q} = π \cdot (d_{q} + c_{r} - l_{c} \cdot \tan α) \cdot \sqrt{{(x_{l a p} - l_{c})}^{2} + {(l_{c} \cdot \tan α + c_{r})}^{2}}

(3)

Relative plunger diameter $d_{a}$ is:

d_{a} = d_{q} - l_{c} \cdot \tan α

(4)

Hydraulic diameter $h_{d}$ is:

h_{d} = \frac{4 \cdot A_{q}}{π \cdot (d_{q} + c_{r} - l_{c} \cdot \tan α)}

(5)

Small orifice flow coefficients $C_{q}$ is:

C_{q} = C_{q \max} \cdot \tanh (\frac{2 λ}{λ_{c r i t}})

(6)

λ = \frac{h_{d}}{υ} \cdot \sqrt{\frac{2 \cdot | p_{q} - p_{z} |}{ρ}}

(7)

λ_{c r i t} \approx \frac{{Re}_{c r i t}}{C_{q \max}}

(8)

where

υ

is the dynamic viscosity of the fluid,

λ_{c r i t}

is the critical flow number, and

{R e}_{c r i t}

is the critical Reynolds number.

The relationship among the valve cylinder flow rate $Q_{q}$ , the valve cylinder oil pressure $p_{q}$ , and the side oil circuit oil pressure $p_{z}$ is

Q_{q} = \frac{d x_{q}}{d t} \cdot \frac{π {d_{q}}^{2}}{4} \cdot \frac{ρ (p_{q})}{ρ (0)} = C_{q} A_{q} \sqrt{\frac{2 \cdot | p_{q} - p_{z} |}{ρ}}

(9)

A force analysis of the valve assembly yields the force balance equation as

\frac{π (d_{q}^{2} - d_{a}^{2})}{4} p_{z} + \frac{π d_{a}^{2}}{4} p_{q} - m_{q} \frac{d^{2} x_{q}}{d t^{2}} - f_{q 0} - k_{q} x_{q} = 0

(10)

When

0 < x_{l a p} \leq x_{2}

, the buffer plunger is still throttled with a small orifice between it and the buffer hole. The flow cross-sectional area

A_{q}

is the side area of the circular table where the vertical segment of FG is located.

A_{q} = π \cdot (d_{q} + c_{r} - x_{l a p} \cdot \cos α \cdot \sin α) \cdot (x_{l a p} \cdot \sin α + \frac{c_{r}}{\cos α})

(11)

Relative plunger diameter $d_{a}$ is:

d_{a} = d_{q} \cdot x_{l a p} \cdot \cos α \cdot \sin α

(12)

Hydraulic diameter $h_{d}$ is:

h_{d} = \frac{4 \cdot A_{q}}{π \cdot (d_{h} + d_{a})}

(13)

Bring $A_{q}$ , $d_{a}$ , and $h_{d}$ into equations (6), (9), and (10) to obtain the small orifice flow coefficient $C_{q}$ , valve cylinder flow equation, and valve assembly force balance equation for this stage.

When $x_{l a p} \leq 0$ , there is circular slit throttling between the buffer plunger and the buffer hole and the flow cross-sectional area $A_{q}$ is:

A_{q} = \frac{π}{4} \cdot (d_{h}^{2} - d_{q}^{2})

(14)

Relative plunger diameter $d_{a}$ is:

d_{a} = d_{q}

(15)

Hydraulic diameter $h_{d}$ is:

h_{d} = \frac{4 \cdot π \cdot (d_{q} + c_{r}) \cdot \frac{c_{r}}{\cos α}}{π \cdot (d_{h} + d_{q})}

(16)

Variation of flow in the valve cylinder $Q_{q}$ is:

Q_{q} = \frac{d x_{q}}{d t} \cdot \frac{π {d_{q}}^{2}}{4} \cdot \frac{ρ (p_{q})}{ρ (0)} = \frac{π \cdot d_{q} \cdot c_{r}^{3}}{12 \cdot ρ \cdot υ} \cdot \frac{| p_{q} - p_{z} |}{k_{g} + x_{l a p}}

(17)

The correction factor $k_{g} = \frac{π \cdot d_{q} \cdot c_{r}^{3}}{12 \cdot ρ \cdot υ} \cdot \frac{| p_{q} - p_{z} |}{Q_{q 0}}$ has been added to ensure continuity of the flow calculation between the small orifice throttle and the slit throttle at $x_{l a p} = 0$ , where $Q_{q 0}$ is the small orifice throttle flow at $x_{l a p} = 0$ obtained using equation (9).

A force analysis of the valve assembly yields the force balance equation as:

\frac{π d_{q}^{2}}{4} p_{q} - m_{q} \frac{d^{2} x_{q}}{d t^{2}} - f_{q 0} - k_{q} x_{q} = 0

(18)

Valve seating simulation modeling

Based on the above mathematical model and the working principle of the system, a simulation model for valve seating was established using AMESim. Figure 8 illustrates this model, and Table 1 outlines the key parameters of the simulation model.

Figure 8.

AMESim numerical simulation model for valve seating.

Table 1.

Main parameters of the simulation model.

No.	Parameter	Unit	Value
1	m_q	g	110
2	f_q0	N	170
3	k_q	N/mm	20
4	d_q	mm	10
5	ρ	g/mm³	870 × 10⁻⁶
6	ξ	—	1.33
7	α	deg	20
8	l_c	mm	4
9	c_r	mm	0.06

The red part is the control subsystem, the blue and brown parts are the hydraulic subsystem, and the green part is the mechanical subsystem. The engine speed is simulated and controlled through the rotational signals of the camshaft. In the valve regulator, the limit rod adjustment function in the physical system is realized by controlling the range of movement displacement of the mass block, which satisfies the adjustability requirement.^28,29

Figure 9 displays a comparison between the test curves and the simulation curves, revealing several observations: the phase of the tested valve closely mirrors that of the simulation; the lift curve of the valve push section demonstrates a notable correspondence with the simulated counterpart; despite a slightly diminished instantaneous lift in the tested valve return section compared to the simulated lift, the overall lift curve alignment is deemed satisfactory; the tested valve lift curve exhibits commendable consistency with the simulated curve; the tested oil pressure curve bears a strong resemblance to the simulated oil pressure curve, featuring peaks and troughs that correspond closely to comparable crank angles and oil pressures. The robust concordance between simulation and experimental results underscores the dependability of the simulation model, thereby facilitating a comprehensive analysis of valve seating characteristics.

Figure 9.

Comparison of test and simulation curves. (a) n = 1000 r/min, A = 0 mm and (b) n = 3000 r/min, A = 1 mm.

Analysis of valve seating characteristics

Regulator adjustment amount

Figure 10 presents the simulation curves of valve velocity for various adjustment amounts at different engine speeds, where the solid red dot in the curve represents the valve seating point. Because the adjustment section of the cam uses the law of equal speed movement, the valve seating velocity corresponding to different adjustment amounts at the same engine speed is basically the same, and the adjustment amount has less influence on the valve seating characteristics. To facilitate the analysis, the adjustment amount is fixed at A = 2 mm in the following.

Figure 10.

Simulation curves of valve velocity at different adjustment amounts. (a) n = 1500 r/min and (b) n = 3500 r/min.

Engine speed

Figures 11 and 12 depict simulation curves representing valve velocity and valve cylinder pressure at various engine speeds. Notably, as the engine speed rises, valve seating velocity experiences a gradual increase, and the valve seating phase exhibits a corresponding lag, indicating a progressive deterioration in valve seating characteristics. Simultaneously, the valve cylinder pressure $p_{q}$ gradually intensifies with escalating engine speed.

Figure 11.

Simulation curves of valve velocity at different engine speeds.

Figure 12.

Simulation curves of valve cylinder oil pressure at different engine speeds.

Buffer mechanism parameters

The conical flange buffer mechanism parameters include the minimum buffer clearance $c_{r}$ , the cone angle $α$ , and the cone length $l_{c}$ . According to the previous analysis, the higher the engine speed, the worse the valve characteristics. Hence, for the purpose of analysis, the engine speed will be fixed at 4500 r/min.

Figures 13 and 14 show the simulation curves of valve velocity and valve cylinder pressure for different minimum buffer clearance $c_{r}$ . It can be seen that the smaller the buffer clearance $c_{r}$ , the smaller the valve seating velocity, but the more lagged the valve seating moment, which affects the accuracy of the valve phase while the valve cylinder pressure also gradually increases.

Figure 13.

Simulation curves of valve velocity at different buffer clearances.

Figure 14.

Simulation curves of valve cylinder oil pressure at different buffer clearances.

Figure 15 illustrates the valve velocity simulation curves at different cone angles. As the cone angle $α$ increases, the valve seating velocity decreases, and the valve seating phase angle lags further. Meanwhile, Figure 16 displays the simulation curves of valve cylinder pressure at different cone angles $α$ . It is evident that changes in the cone angle $α$ exert a more pronounced influence on the valve cylinder pressure. Specifically, when $α = 40 °$ , the valve cylinder oil pressure approaches 20 MPa.

Figure 15.

Simulation curves of valve velocity at different cone angles.

Figure 16.

Simulation curves of valve cylinder oil pressure at different cone angles.

Figures 17 and 18 show the simulation curves of valve velocity and valve cylinder pressure for different minimum buffer lengths $l_{c}$ , respectively. It can be seen that the variation in buffer length has a small effect on valve movement velocity and valve cylinder pressure.

Figure 17.

Simulation curves of valve velocity for different buffer lengths.

Figure 18.

Simulation curves of valve cylinder pressure at different buffer lengths.

Optimization of valve seating characteristics

An optimization model was established by Design Exploration in order to ensure the stability and phase accuracy of the valve seating. Additionally, the genetic algorithm was chosen for the optimization process, and the specific parameters are shown in Table 2.

Table 2.

Specific parameters of the genetic algorithm.

No.	Parameter	Value
1	Population size	50
2	Reproduction ratio	80%
3	Max. number of generation	20
4	Mutation probability	5
5	Mutation amplitude	0.2

In the previous analysis, it was found that the buffer clearance $c_{r}$ , cone angle $α$ , and buffer length $l_{c}$ all affect the seating performance of the valve. Therefore, variables $c_{r}$ , $α$ , and $l_{c}$ were selected as the optimization parameters. The working point was set at the maximum system speed of 4500 r/min and adjustment amount is fixed at A = 2 mm. The objective was to reduce the phase delay of the valve seating as much as possible while meeting the requirements for valve seating velocity and allowable buffer chamber pressure.³⁰ When the valve is seated, the impact force will be generated with the valve seating, and the acceleration of the valve motion will have a large mutation, so the seating phase can be determined when the valve acceleration has a large mutation.

(1) Definition of objective function

The minimum value of valve seating phase angle $\min β$ is:

\min β = f (c_{r}, α, l_{c})

(19)

(2) Definition of optimization variable parameters

The range of buffer clearance $c_{r}$ variation is:

0.01 m m \leq c_{r} \leq 0.06 m m

(20)

The range of cone angle $α$ variation is:

5^{\circ} \leq α \leq 45^{\circ}

(21)

The range of buffer length $l_{c}$ variation is:

1 m m \leq l_{c} \leq 5 m m

(22)

(3) Definition of constraint conditions

The constraint value for valve seating velocity $v_{q s}$ is:

v_{q s} \leq 0.5 m / s

(23)

The constraint value for buffer chamber pressure $p_{q}$ is:

p_{q} \leq 17 M p a

(24)

(4) Optimization process and results

As shown in Figures 19–22, with variables varying within the valid range, after several hundred iterations, the optimization objective tends to converge. The final optimized results are presented in Table 3.

(5) Test analysis

Figure 19.

Curve of the variation process of buffer clearance.

Figure 20.

Curve of the variation process of buffer cone angle.

Figure 21.

Curve of the variation process of buffer length.

Figure 22.

Curve of the optimization process of the objective.

Table 3.

Optimized results.

No.	Parameter	Unit	Value
1	Buffer clearance $c_{r}$	mm	0.025
2	Buffer cone angle $α$	deg	12.441
3	Buffer length $l_{c}$	mm	3.809
4	Valve seating phase angle $\min β$	deg	589.688

Based on the optimization results in Table 3, the values for the buffer mechanism parameters are determined, as indicated in Table 4. The buffer mechanism is reprocessed, and the valve seating performance test is carried out using the physical test platform. The test results are shown in Figure 23 and Table 5.

Table 4.

Test parameters for the buffer mechanism.

No.	Parameter	Unit	Value (Initial)	Value (Current)
1	Buffer clearance $c_{r}$	mm	0.06	0.025
2	Buffer cone angle $α$	deg	20	12.5
3	Buffer length $l_{c}$	mm	4	3.8

Figure 23.

Valve movement pattern test curves at an engine speed of 4500 r/min.

Table 5.

Valve seating performance comparison (n = 4500 r/min, A = 2 mm).

No.	Parameter	Unit	Value (Initial)	Value (Current)
1	Valve seating velocity	m/s	1.33	0.52
2	Valve seating rebound height	mm	0.32	0.08
3	Valve seating phase angle	deg	588.2	591.5
4	Maximum buffer chamber pressure	MPa	16.4	17.2

It can be seen that the current valve seating velocity is reduced from 1.33 m/s to 0.52 m/s, and the rebound height is reduced from 0.32 mm to 0.08 mm, which meets the valve seating velocity requirements. Although there is a delay in the valve seating phase, the delay angle is only 3.3°, and the corresponding maximum pressure rise of the buffer chamber is only 0.8 MPa, which can still meet the system pressure requirements.

Conclusions

Ensuring smooth valve seating is a key technology for hydraulic variable valve systems. Therefore, we established mathematical and simulation models for the valve seating process, optimized the parameters of the hydraulic buffer mechanism, and conducted physical platform testing experiments and analyses. The following conclusions were obtained:

(1) In hydraulic variable valve systems, the valve seating characteristics not only ensure smooth valve seating but also take into account the timeliness of the valve seating and the permissible valve cylinder pressure requirements. The model of the buffer mechanism for valve seating has proven to be highly reliable. Simulation results depicting valve movement align well with corresponding test results. This model is applicable for the comprehensive analysis of valve seating characteristics and their influencing factors.

(2) The valve seating characteristics were analyzed with different adjustment amounts, different engine speeds, and different buffer mechanism parameters. The engine speed, the minimum buffer clearance $c_{r}$ , and the cone angle $α$ had a significant influence on the valve seating velocity, seating phase, and valve cylinder pressure. The higher the engine speed, the worse the valve seating characteristics. The smaller the buffer clearance, the smaller the valve seating velocity, but the more lagged the valve seating moment, while the valve cylinder pressure also gradually increases. As the cone angle increases, the valve seating velocity decreases, and the valve seating phase angle lags further.

(3) Selecting the system working speed of 4500 r/min with the worst performing case as the research point, optimized test results reveal that, while maintaining compliance with allowable oil pressure, the valve seating velocity has been significantly reduced from 1.33 m/s to 0.52 m/s. The seating phase delay is merely 3.3°, showcasing smooth valve seating and effectively expanding the application range of the variable valve system.

(4) In this paper, the control system of a single intake valve is studied. Next, the multi-valve integrated system and variable exhaust valve system will be studied to promote the commercialization of this technology as soon as possible.

Statements and declarations

Footnotes

Declaration of conflicting interests

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

Funding

The author disclosed receipt of the following financial support for the research, authorship, and/or publication of this article: This work was funded by Guizhou Provincial Basic Research Program (Natural Science) (Grant No.: Qianke-hejichu-[2020]1Y226), Guiyang University Multidisciplinary Team Construction Projects in 2021[2021-xk09], and Introducing Talents to Initiate Funded Research Projects of Guiyang University (Grant No.: GYU-ZRD[2018]-027).

ORCID iD

Yuliang Xu

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