Design and Experiment of Self-Propelled High-Stem Chrysanthemum coronarium Orderly Harvester

Abstract

To address the issues of low efficiency, high cost of manual harvesting, and the lack of mechanized harvesting technology and equipment for high-stem Chrysanthemum coronarium, a self-propelled orderly harvester was designed to perform key harvesting operations such as row alignment, clamping and cutting, orderly conveying, and collection. Based on the analysis of agronomic requirements for cultivation and mechanized harvesting needs, the overall structure and working principle of the machine were described. Meanwhile, the key components such as the reciprocating cutting mechanism and orderly conveying mechanism were structurally designed and theoretically analyzed. The main structural and operating parameters of the harvester were determined based on the geometric and kinematic conditions of high-stem Chrysanthemum coronarium during its movement along the conveying path, as well as the mechanical model of the conveying process. In addition, a three-factor, three-level Box-Behnken field experiment was also conducted with the experimental factors including the machine’s forward, cutting, and conveying speed, and evaluation indicators like harvesting loss rate and orderliness. A second-order polynomial regression model was established to analyze the relationship between the evaluation indicators and the factors using the Design-Expert 13 software, which revealed the influence patterns of the machine’s forward speed, reciprocating cutter cutting speed, conveying device speed, and their interaction influence on the evaluation indicators. Moreover, the optimal parameter combination, obtained by solving the optimization model for harvesting loss rate and orderliness, was forward speed of 260 mm/s, cutting speed of 250 mm/s, and conveying speed of 300 mm/s. Field test results showed that the average harvesting loss rate of the prototype was 4.45% and the orderliness was 92.57%, with a relative error of less than 5% compared to the predicted values. The key components of the harvester operated stably, and the machine was capable of performing cutting, orderly conveying, and collection in a single pass. All performance indicators met the mechanized orderly harvesting requirements of high-stem Chrysanthemum coronarium.

1. Introduction

Leafy vegetables are a type of fast-growing vegetable primarily consumed for their green leaves and tender stems, including high-stem Tung Ho (Chrysanthemum coronarium), Pak choi (Brassica rapa subsp. chinensis), spinach (Spinacia oleracea), and celery (Apium graveolens). Due to their short growth cycle and rich nutritional value, leafy vegetables are widely cultivated in China, accounting for 37.15% of the total vegetable planting area and 30.5% of total production. High-stem Chrysanthemum coronarium, a member of the Asteraceae family, is characterized by dark green leaves, small and thin leaf blades, pale green slender stems, a plant height of 30–40 cm, low fiber content, high quality, strong adaptability, and fast growth [1]. It is suitable for cultivation throughout China, where the yield has reached 45,000 kg/hm². From sowing to harvest, its typical growth cycle lasts for 30 to 40 days [2]. The stems and leaves of high-stem Chrysanthemum coronarium are high in relative water content (RWC) and very tender, which has led to the fact that there has been no suitable mechanized harvesting equipment so far, which has finally resulted in predominantly manual harvesting. This leads to high labor intensity, low efficiency, and high production costs, which constrain the efficient development of the vegetable industry [3].

According to the degree of orderliness during the collection process, leafy vegetable harvesters can be categorized into unordered harvesters and orderly harvesters. Unordered harvesters are compact in structure, highly adaptable, and reliable. However, the harvested vegetables are often disorganized, requiring manual post-sorting [4,5,6]. In developed countries, the technology for harvesting stem-and-leaf vegetables has been well established. The SLIDE CRAB leafy vegetable harvester developed by Italy’s Hortech company uses a flexible dual-corrugated clamping conveyor system to transfer vegetables to a manual operation platform, which could reduce the damage to plants during conveying. Besides, another model, the SLIDE VALERIANA ECO, employs a reciprocating cutting mechanism to cut vegetable roots and a vibrating screen to remove debris from the leaves. It then conveys the crop rearward via a rotary belt to a transverse wave-shaped conveyor, which pushes vegetables into side-mounted collection boxes [7]. Italy’s Ortomec developed the 8400 ELECTRA, a high-automation self-propelled leafy vegetable harvester, which features automatic driving and cutting platform height control. That is, it uses a photoelectric sensor to measure the plant growth height in real time and controls a hydraulic cylinder to adjust the distance between the cutter and the ground, ensuring consistent stubble height [8]. To meet agricultural production needs, Japan and South Korea have also developed harvesters suitable for their own leafy vegetable cultivation models. Japan’s Kawasaki company developed a walk-behind leafy vegetable harvester with a simple structure that uses high-speed airflow to deliver cut vegetables into a collection bag. Korea’s Plan T company developed the MT-200 walk-behind leafy vegetable harvester, known for its reliability and ease of operation, which is suitable for small-scale farmers in Asia and is used for crops like spinach and rape under precision farming systems [9]. Although China leads the world in vegetable planting area and yield, research and development of vegetable harvesters began relatively late, and only a few models are currently available [10,11]. Most vegetable harvester technologies are still in the early development stage. Extensive research has been conducted on the design and optimization of key components such as the cutting and conveying mechanisms of harvesters. In terms of crop cutting, the primary cutting methods include dual-disc cutters, single-disc cutters, and reciprocating cutters [12,13,14,15,16,17,18,19,20,21]. For crop conveying, commonly used methods include belt-type conveying, synchronous belt clamping, and flexible clamping conveying systems [22,23,24,25,26,27,28]. In the field of orderly harvesting technology and equipment development, Liu Dong studied the mechanical properties of Pak choi and developed a walk-behind orderly harvester for it, using a twisted clamping conveyor to achieve orderly conveying [29]. Song Ye constructed a low-damage constraint model for spinach, analyzed forces during pulling and conveying, and optimized key parameters of the pulling mechanism based on the results [30]. Zou Liangliang [31] proposed a theoretical model of the rheological properties of spinach during clamping and conveying. Based on the Burgers viscoelastic model, he built a rheological constitutive equation and determined the relationship between the working parameters of the clamping conveyor and the minimum plastic deformation of spinach through experimental analysis. Gao Feng and others from Jiangsu University conducted in-depth studies on key components such as the cutting and clamping conveying mechanisms of orderly harvesters for leafy vegetables, and optimized the working parameters of the conveyor system based on the damage rate of baby bok choy [32]. Shi Yinyan [33] and colleagues from Nanjing Agricultural University developed a pure battery-driven self-propelled orderly harvesting system for leafy vegetables. During operation, the vegetables are vertically clamped by a clamping conveyor mechanism, then transferred to a top-end steering device. After the bottom of the vegetables is intercepted, they fall onto a horizontal conveyor belt and are delivered to a side-mounted collection box, realizing orderly harvesting of leafy vegetables [33]. Liao Qingxi [34] et al. designed a self-propelled oilseed rape stem and leaf combined harvesting machine, which can complete processes such as cutting, unorganized transportation, and collection. The optimal working parameters of the entire machine were determined through theoretical analysis and bench tests. Liao Yitao and others [35] designed a gantry-type electric-driven rapeseed stalk harvester, which uses a gantry crawler chassis and a vertical rotary chopping mechanism. Optimal structure and operation parameters were determined through theoretical analysis and bench tests, achieving mechanized harvesting of rapeseed stalks [35]. Li Hai et al. developed a row-oriented orderly harvester for rapeseed stalks, which achieved orderly laying and collection of stalks through a combination of twisted conveying and diversion mechanisms [36]. Liao Yitao et al. designed a self-propelled row-oriented rapeseed stalk harvester that adopted a working sequence of bidirectional cutting, clamping conveying, lateral conveying, and collection box integration to realize orderly harvesting of rapeseed stalks [37]. At present, there is limited research on harvesters specifically designed for high-stem Chrysanthemum coronarium, and the aforementioned vegetable harvesters are insufficient to meet the agronomic and mechanized harvesting requirements of high-stem Chrysanthemum coronarium. At present, there is relatively little research on harvesting machinery for high-stem Chrysanthemum coronarium. Compared with spinach and rape sprouts, high-stem Chrysanthemum coronarium are consumed for their tender leaves and stems, which have the physical characteristic of being soft. The existing harvesting machinery exerts squeezing and friction on high-stem Chrysanthemum coronarium, causing the tender leaves and stems to be easily damaged. To meet the requirements of efficient cutting, low-damage clamping, and orderly conveying, the above-mentioned vegetable harvesters are difficult to meet the agricultural mechanization harvesting requirements of high-stem Chrysanthemum coronarium.

To meet the demands for efficient, low-damage, and orderly harvesting of high-stem Chrysanthemum coronarium and improve the mechanization level of its harvesting process, this study proposes an orderly harvesting strategy based on the agronomic requirements of cultivation and harvesting. The strategy includes “row-alignment separation + dual-action cutting + synchronized bottom and clamping conveying + flow-guided collection”. A Self-propelled high-stem Chrysanthemum coronarium orderly harvester was designed, which mainly consists of a self-propelled electric chassis, a row separator, a reciprocating cutting mechanism, a cutting platform height adjustment system, a flexible clamping conveyor, and guiding plates. Through theoretical analysis and kinematic simulation, the structures and operating parameters of key working components were determined. Taking into account the test costs and the effective collection and analysis of experimental samples, the Box-Behnken design is time- and cost-efficient, making it easier to collect and analyze the samples. Therefore, we used the loss rate of harvest and the degree of orderliness as the evaluation indicators. Through conducting the Box-Behnken design, we established a regression model between the evaluation indicators and the experimental parameters. An optimal parameter combination for the orderly harvester was obtained and verified through field tests, aiming to provide a reference for the development of mechanized and orderly harvesting equipment for high-stem Chrysanthemum coronarium.

2. Materials and Methods

2.1. Materials and Test Equipment

2.1.1. Overall Structure and Working Process

The mechanized cultivation of high-stem Chrysanthemum coronarium is primarily based on mechanical direct seeding. The schematic diagram of agronomic requirements for planting and harvesting is shown in Figure 1. The ridge top, furrow, ridge bottom width, ridge height, and planting spacing, respectively, were 850–900 mm, 200–250 mm, 1000–1200 mm, 200–250 mm, and 20–25 mm. During the harvesting period, the plant height reached 300–400 mm, and the stubble height for mechanized harvesting was 20–30 mm.

High-stem Chrysanthemum coronarium exhibits the following morphological characteristics: the plant height (L) ranges from 30 to 40 mm; the leaf spread width (D) measures between 10 and 15 mm; the stem diameter (d) at 2 cm from the node falls within 3–6 mm; the leaf blade length (L₁) is 15–23 mm; and the primary root length (L₂) spans 3–6 mm. The size data of the Chrysanthemum coronarium was shown in Figure 2.

The self-propelled high-stem Chrysanthemum coronarium orderly harvester is mainly composed of a divider, a reciprocating electric cutting mechanism, a header height adjustment mechanism, an orderly conveying mechanism, a guiding mechanism, a floating collection mechanism, a collection platform, and a self-propelled electric chassis, as shown in Figure 3.

The self-propelled electric four-wheel chassis mainly includes a frame, power system, drive system, and control system. The chassis travels within the furrows to reduce the trampling damage to the crops during operation. The orderly harvesting header is composed of operational components such as the divider, reciprocating cutting mechanism, vertical conveyor belt, bottom rotary conveyor belt, guide plate, floating collection mechanism, and drive motors, all mounted on the frame.

The header height adjustment mechanism utilizes symmetrically arranged electric actuators to adjust the working height and inclination of the orderly conveying header, thereby improving adaptability and passability during harvesting. The divider, installed at the front end of the orderly conveying header, separates the plant rows of high-stem Chrysanthemum coronarium, while also gathering and guiding the plants for in-row harvesting. The orderly conveying mechanism consists of a vertical clamping conveyor module, a rotary bottom conveyor module, and a drive module, which realize the orderly transport of the crop. The reciprocating cutting mechanism is arranged at the lower front side of the conveying device. The floating collection mechanism is hinged to the harvesting header and can rotate around the hinge point to adjust the collection height according to different working conditions as the header height changes.

During harvesting, as the harvester moves forward, the divider separates and gathers the crop into rows and guides it to the front end of the inclined orderly conveying mechanism, where contact with the conveyor belts is established. Once the stem is effectively clamped by the conveyor, the reciprocating cutter mounted at the lower front side of the conveying mechanism provides supported cutting of the stem. The cut crop is then conveyed diagonally upward by the orderly conveying mechanism to the end of the clamping conveyor. The guide plate directs the harvested Chrysanthemum coronarium vertically and orderly into the collection basket on the floating collection mechanism. Once the basket is full, it is unloaded onto the collection platform, completing the orderly harvesting process. During operation, the harvesting height can be adjusted via electric actuators, and both the reciprocating cutting speed and conveyor speed are steplessly adjustable to meet harvesting requirements under different crop conditions and working environments. The main technical parameters of the self-propelled high-stem Chrysanthemum coronarium orderly harvester are shown in

Table 1. Main technical parameters.

Parameter	Value/Type
Traveling system	Self-propelled electric four-wheel chassis
Overall dimensions (L × W × H)/(mm × mm × mm)	3000 × 1600 × 1500
Weight/kg	600
Drive mode of working components	Electric
Power/kW	2.0
Operating speed/(m·s−1)	0.2–0.5
Number of rows harvested	3
Stubble height/mm	0–500
Cutting width/mm	≤900
Harvesting efficiency/(hm2·h−1)	≥0.1
Harvesting loss rate/%	≤5%
Orderliness/%	≥90%

.

2.1.2. Design of the Reciprocating Electric Cutting Mechanism

The reciprocating electric cutting mechanism is installed at the front end of the cutter platform of the self-propelled high-stem Chrysanthemum coronarium orderly harvester. It mainly consists of a driving mechanism and blades. The driving mechanism includes a motor, gears, double eccentric cams, and crank arms. The structural layout is shown in Figure 4.

The mechanism is independently driven by the motor, which transmits power through a pair of transmission gears to drive the double eccentric cams. Crank arms are mounted on the eccentric cams, and the ends of the crank arms are connected to the upper and lower blades via fixed pins. The motor drives the upper and lower blades to perform a reciprocating motion through the transmission mechanism, thereby achieving a double-action cutting effect.

To achieve symmetrical double-action cutting in opposite phase, a symmetrical dual-eccentric wheel driving mechanism was adopted, as shown in Figure 5. Let O₁ and O₂ represent the centers of the first and second eccentric wheels, respectively, with an eccentricity of e. The eccentric shaft is fixed at the midpoint of the line segment connecting O₁ and O₂. The two eccentric wheels rotate synchronously, and a coordinate system is established with the eccentric shaft center O as the origin. The crank arms connect the eccentric wheels and blades, with points A and B denoting the connecting points of the crank arms to the two blades. Both crank arms O₁A and O₂B have a length of l. The blades move along the x-axis, and each point on the blades shares the same motion pattern in the x-direction. The displacement of the driving mechanism can be expressed based on geometric relationships. The vertical coordinates of points A and B are:

$$\left\{\begin{array}{l}{x}_A=l\cos {\theta }_2+e\cos {\theta }_1 {\theta }_1\subset \left({\displaystyle \frac{\pi }{2}},{\displaystyle \frac{3\pi }{2}}\right)\\ x_B=l\cos {\theta }_2-e\cos {\theta }_1 {\theta }_1\subset \left(-{\displaystyle \frac{\pi }{2}},{\displaystyle \frac{\pi }{2}}\right)\end{array}\right.$$

(1)

where x_A, x_B are the horizontal coordinates of points A and B, respectively; θ₁ is the angle between the line O₁O₂ and the x-axis (°); and θ₂ is the angle between the crank arm and the x-axis (°).

According to Equation (1), the motion trajectories of the upper and lower blades are identical but have a phase difference of π. At any moment, the displacements of the upper and lower blades from their initial positions are equal in magnitude and opposite in direction. When ${\theta }_1=n\pi (n=0,1,2\cdots ),{\theta }_2=0$, the displacement reaches its peak, its magnitude equals the eccentricity e.

The speed of the reciprocating electric cutting device is analyzed. Differentiating Equation (1) with respect to time, the expression for the cutting speeds of the upper and lower blades is obtained as:

$$\left\{\begin{array}{l}{\upsilon }_A=-l\sin {\theta }_2·{\omega }_1-e\sin {\theta }_1·{\omega }_2 {\theta }_1\subset \left({\displaystyle \frac{\pi }{2}},{\displaystyle \frac{3\pi }{2}}\right)\\ {\upsilon }_B=-l\sin {\theta }_2·{\omega }_1+e\sin {\theta }_1·{\omega }_2 {\theta }_1\subset \left(-{\displaystyle \frac{\pi }{2}},{\displaystyle \frac{\pi }{2}}\right)\end{array}\right.$$

(2)

where v_A, v_B are the velocities at points A and B (m/s); ω₁ is the angular velocity of the crank (rad/s); and ω₂ is the angular velocity of the eccentric wheel (rad/s).

According to Equation (2), the velocity curves of the upper and lower blades are identical in shape and opposite in phase. At any moment, the upper and lower blades have equal speeds but in opposite directions. When ${\theta }_1=n\pi /2(n=0,1,2\cdots ),\sin {\theta }_2=e/l$, the cutting speed of the blade reaches its maximum value of $e\left|{\omega }_1-{\omega }_2\right|$.

The accelerated speed of the reciprocating electric cutting device is analyzed. Differentiating Equation (2) with respect to time, the expression for the cutting speeds of the upper and lower blades is obtained as:

$$\left\{\begin{array}{l}{a}_A=l{\omega }_1^2\cos {\theta }_2+l\sin {\theta }_2{\alpha }_1+e{\omega }_2^2\cos {\theta }_1+e\sin {\theta }_1{\alpha }_2\hspace{1em}\theta \subset \left(-{\displaystyle \frac{\pi }{2}},{\displaystyle \frac{3\pi }{2}}\right)\\ a_B=-l{\omega }_1^2\cos {\theta }_2+l\sin {\theta }_2{\alpha }_1+e{\omega }_2^2\cos {\theta }_1+e\sin {\theta }_1{\alpha }_2 \theta \subset \left(-{\displaystyle \frac{\pi }{2}},{\displaystyle \frac{\pi }{2}}\right)\end{array}\right.$$

(3)

where a_A, a_B are the accelerations at points A and B (m/s²); and α₁, α₂ are the angular accelerations of the crank arms O₁A and O₂B (rad/s²).

According to Equation (3), the acceleration curves of the upper and lower blades are also identical in shape and opposite in phase. At any time, the magnitudes of their accelerations are equal and directions are opposite, which results in the mutual cancellation of inertial forces, thereby reducing the vibration of the cutter platform. When ${\theta }_1=n\pi (n=0,1,2\cdots ),{\theta }_2=0$, the cutting speed of the blade reaches its maximum value of $\left|e{\omega }_2^2-l{\omega }_1^2\right|$.

According to the design principle of involute cylindrical gears, to avoid undercutting, the number of gear teeth should exceed the minimum limit, and the gear pair should have coprime tooth counts. The driving gear is designed with 14 teeth, and the driven gear with 58 teeth, resulting in a reduction ratio of 4:1. The gear module is 1.25, the pressure angle is 20°, and the root circle diameter of the large gear is 70 mm. From Equations (1)–(3), it is evident that the key parameters affecting the kinematic performance of the reciprocating electric cutting mechanism are the eccentricity e and crank length l. The eccentricity e determines the motion range of the blades and affects the gripping and cutting performance on plant stems. The crank length l influences the stroke-to-speed ratio of the blade movement. Considering the root circle diameter of the large gear, the eccentricity and crank length are designed as e = 8.5 mm and l = 57 mm, respectively.

The moving blade is the primary working component of the reciprocating cutting mechanism. It adopts a smooth-edge design, which facilitates efficient cutting and ensures a clean stubble. The blade is made of 65Mn spring steel. The main structural parameters of the moving blade include the cutting angle α (i.e., the inclination of the blade edge), the blade edge height h, and the blade widths C and d. The cutting angle α is a critical design parameter of the moving blade. It significantly affects the cutting performance, including the cutting resistance, and determines whether the blade can securely grip the stem of the vegetable to ensure reliable cutting. As shown in Figure 6, C represents the rear width of the blade, d is the front width, h is the height of the cutting edge, A is the starting point of the cutting edge, α is the cutting angle, and v is the motion velocity of the blade. Then

$${\upsilon }_1=\upsilon \sin \alpha$$

(4)

During cutting, to ensure the vegetable stem is gripped securely, the resultant forces R₁ and R₂ acting from the two cutting edges must be collinear. A schematic illustration of the blade gripping the vegetable stem is shown in Figure 7. In the figure, ${\phi }_1={\phi }_2$ represents the friction angle of the moving blade on the vegetable stem, and α is the cutting angle; R₁ and R₂ are the resultant forces exerted by the two cutting edges on the stem.

According to Figure 6, it can be seen that:

$$F_1\le N_1\tan {\phi }_1$$

(5)

$$F_2\le N_2\tan {\phi }_2$$

(6)

$$\left\{\begin{array}{l}{F}_1=F_2\\ {\phi }_1={\phi }_2\end{array}\right.$$

(7)

The condition for gripping the vegetable stem securely is: $2\alpha \le {\phi }_1+{\phi }_2$. Since the geometric and motion parameters of the upper and lower sets of shears are the same, then α ≤ φ. Based on measured data, the friction angle between the blade and the stem ranges from 20° to 25°; thus, this paper selects a cutting angle α = 19°, which satisfies the stable gripping requirement. The blade edge height h affects the load per unit blade length, the cutting profile, and the amount of vertical inclination. The relationship between h and the cutting angle α can be expressed as:

$$h={\displaystyle \frac{C-d}{2\tan \alpha }}$$

(8)

Based on design calculations and requirements, the selected parameters are: Rear blade width C = 14 mm. Front blade width d = 7 mm. Blade edge height h = 20 mm. Blade center distance b = 2e = 17 mm. Total blade length L = 1000 mm. The blade is manufactured as a single integrated piece to reduce vibration during operation. The detailed engineering drawing of the reciprocating cutting blade is shown in Figure 8.

To verify the theoretical calculation results of the reciprocating cutting mechanism, the Adams motion simulation software (v2022) was used to conduct a kinematic simulation analysis of the cutting device under no-load conditions. First, a 3D model of the cutting device was built in SolidWorks (v2022). To reduce computational load, the motor and connecting components were simplified, retaining only the essential parts such as the output gear, double eccentric spindle, cutter, and base. The model was then saved in Parasolid (*.x_t) format and imported into Adams (v2022). For ease of adding constraints and drives, all components were uniformly named and numbered, and constraints were applied accordingly.

To simulate the actual working conditions of the cutting device, motion drives were added. A rotational drive was applied to the revolute pair between the bearing and the double eccentric spindle, and a translational drive was applied to the prismatic pair between the base and the ground. After all constraints and drives were added, the setup was shown in Figure 9. After parameter settings were completed, the simulation was initiated with a termination time of 3 s and 50 steps.

After the simulation, the Postprocessor function was used to plot the motion curves of the cutter. The displacement–time, velocity–time, and acceleration–time curves of the upper and lower cutters are shown in Figure 9 and Figure 10, respectively.

When the cutter was at the origin, the cutting speed was 0, and the acceleration reached its maximum. When the output gear rotated 90°, the cutter was at the midpoint of the cutting stroke, the cutting speed reached its maximum, and the acceleration dropped to 0—this was the cutting stage. When the output gear rotated 180°, the cutter reached the bottom of the stroke, the speed returned to 0, and the acceleration increased again to the maximum. When the output gear rotated 270°, the cutter was at the midpoint of the return stroke, the cutting speed was again at its maximum, and the acceleration returned to 0—this corresponded to the cutter cutting vegetables during the return stroke. When the output gear rotated 360°, the cutter completed a full reciprocating motion and returned to the origin, with cutting speed back to 0 and acceleration again at its maximum.

From Figure 10, it can be seen that the upper cutter had a maximum displacement of 0.414 m and a minimum of 0.397 m, resulting in a cutting stroke of 0.017 m. The maximum positive x-axis velocity was 0.487 m/s, and the maximum negative x-axis velocity was −0.478 m/s. The maximum positive x-axis acceleration was 23.9 m/s², and the maximum negative x-axis acceleration was −32.4 m/s².

From Figure 11, the lower cutter had a maximum displacement of 0.403 m and a minimum of 0.394 m, with a cutting stroke of 0.017 m. The maximum positive x-axis velocity was 0.487 m/s, and the maximum negative x-axis velocity was 0.486 m/s. The maximum positive x-axis acceleration was 23.8 m/s², and the maximum negative x-axis acceleration was −33.6 m/s². The simulation results were consistent with the analytical results, and the displacement, velocity, and acceleration of the upper and lower cutters showed good symmetry, which was beneficial for balancing the inertial forces of both cutters. This indicated that the design of the mechanism was reasonable.

2.1.3. Design of the Orderly Conveying Mechanism

The orderly conveying platform is the core component of the self-propelled high-stem Chrysanthemum coronarium orderly harvester. Its structure was shown in Figure 12. The main components of the orderly conveying mechanism include the divider, vertical clamping conveying module, rotary bottom conveying module, and drive module. Among them, the vertical clamping conveying module and rotary bottom conveying module mainly consist of driving rollers, driven rollers, vertical conveyor belts, rotary bottom conveyor belts, and tensioning mechanisms. The drive module mainly consists of a DC motor, driving sprocket, driven sprocket, bevel gear transmission box, and synchronous pulleys. The drive module is connected to the vertical clamping conveying module and the bottom conveying module through support side plates and the machine frame.

During operation, the DC motor drove the driving sprocket, which transmitted power to the driven sprocket through a chain, thereby driving the horizontal drive rollers to rotate and realizing the rotary motion of the bottom conveyor belt. Through the bevel gear set, the power direction was changed, driving the synchronous pulleys to rotate at the same speed. This drove the left and right vertical drive rollers to rotate in opposite directions, thereby driving the corresponding vertical clamping conveyor belts and rotary bottom conveyor belts to form an upward-slanted conveying path. After being cut by the reciprocating cutting mechanism, the high-stem Chrysanthemum coronarium was orderly conveyed upward along the inclined conveying path and guided into the collection basket one by one through the diversion device.

Clamping the high-stem Chrysanthemum coronarium plants during cutting was essential for achieving orderly conveying. During operation, the orderly conveying mechanism effectively restricted the degree of freedom of the plants perpendicular to the harvester’s traveling direction [38,39]. However, under the influence of factors such as the forward speed of the machine, conveying speed, and conveying angle, the plants might have leaned forward or backward. To ensure effective conveying of the plants by the mechanism, it was necessary to reasonably design the structural form and determine key operating parameters. The posture of the plants during the pre-cut clamping and clamping-cutting stages was analyzed as follows.

(1)

Pre-Cut Clamping Stage

As the machine moved forward, the divider gathered the high-stem Chrysanthemum coronarium plants and guided them into the conveying path formed by the bottom rotary belt and vertical conveying belt. The orderly conveying mechanism clamped the plants effectively. Assume the clamping point was E. The kinematic analysis of the plant posture during this stage is shown in Figure 13. According to the geometric relationship in Figure 9, the movement velocity of the belt was the vector sum of the harvester’s forward speed v_m, the vertical conveying belt speed v_l, and the rotary belt speed v_h.

To avoid jamming during conveying, the velocity of the belts in the harvester’s forward direction must exceed the machine’s forward speed. Thus, the velocities must satisfy:

$$\left({\upsilon }_l+{\upsilon }_h\right)\cos \beta >{\upsilon }_{\mathrm{m}\max }$$

(9)

To ensure the plant remains stationary relative to the clamping device during conveying:

$${\upsilon }_l={\upsilon }_h$$

(10)

The speed of the vertical conveying belt is given by:

$${\upsilon }_l={\upsilon }_h=n_z\pi d_z/60$$

(11)

where v_mmax—maximum forward speed of the harvester,

β—conveying angle with respect to the horizontal plane (°),

n_z—rotational speed of the driving roller, r/min,

d_z—maximum diameter of the conveyor roller.

To maintain an ideal conveying posture, the plant’s vertical speed variation must be minimized. An excessive or insufficient conveying angle β will affect the orderly conveying performance. Therefore, the conveying angle should be within the range of 20° to 30°. Given the forward speed v_m of 0.2–0.5 m/s, and according to Equations (9) and (10), the required vertical and rotary belt speeds v_l and v_h should be within 0.27–0.65 m/s.

(2)

Clamping and Cutting Stage

As the machine moves forward, the cutter contacts the clamped stem and performs a supported cutting. To prevent the plant from failing to enter the clamping track after cutting and to properly design the conveying mechanism, posture variation of the plant in the machine’s traveling direction is analyzed.

As shown in Figure 14, the contact point O between the plant and the ground during the uncut clamping stage is selected as the origin of a Cartesian coordinate system. Assuming the machine is stationary relative to the ground, the plant moves at speed v_m relative to the harvester. The geometric relationship is shown in Figure 10. Let the inclination angles between the plant and the ground during clamping and cutting be γ₁ and γ₂, the conveying angle be β, cutter height from the ground be h_g, cutter position s (horizontal distance between the clamping point and cutter), clamping point height be h_j, and time from start of clamping to cutting be t.

Since the plant’s motion results from the combination of the forward speed v_m, vertical conveying belt speed v_l, and rotary belt speed v_h, the displacement of the clamping point F’ relative to F during this stage is given by:

$$\left\{\begin{array}{l}{x}_{F’}=\left(v_l+v_h\right)t\sin \beta \\ y_{F’}=\left(v_l+v_h\right)t\cos \beta \end{array}\right.$$

(12)

As shown in Figure 13, the inclination angle γ of the plant during the clamping-cutting stage satisfies the trigonometric relationships in triangles △BCH’ and △BEF’:

$$\tan {\gamma }_2={\displaystyle \frac{h_g}{s+h_j\cot {\gamma }_1-v_{m}t}}$$

(13)

$$\tan {\gamma }_2={\displaystyle \frac{\left(v_l+v_h\right)t\sin \beta +h_j}{h_j\cot {\gamma }_1+\left(\left(v_l+v_h\right)\cos \beta -v_m\right)t}}$$

(14)

From Equations (10) and (11):

$${\gamma }_2=\arctan {\displaystyle \frac{\left(v_l+v_h\right)t\sin \beta +h_j-h_g}{\left(v_l+v_h\right)\cos \beta -s}}$$

(15)

Equation (15) indicates that the plant’s posture variation is influenced by parameters such as v_l, v_h, h_g, s, and β, which can be adjusted using the telescopic mechanism of the support frame to meet operational needs.

During orderly harvesting, the clamping center of the conveying belt must be located below the center of mass of the Chrysanthemum coronarium stem to ensure stable conveying and prevent tipping. The theoretical force analysis of the clamped stem is shown in Figure 15. Suppose the cut length of the stem is L, and the harvester’s forward speed and conveyor speeds remain constant, the forces in the vertical and horizontal directions are in equilibrium.

Based on the force analysis in Figure 14, the following equation is established:

$$\left\{\begin{array}{l}{F}_l\sin \beta +F_h\cos \beta =G\\ F_t+F_h\sin \beta =F_l\cos \beta \\ F_l=\mu N\end{array}\right.$$

(16)

where G—weight of the plant (N),

F_t, F_l—force components of the plant weight (N),

β—conveying angle (°),

N—clamping force of the vertical conveying belt (N),

μ—friction coefficient between clamping belt and plant.

Given the plant mass, friction coefficient, and the angle β between the conveying structure and the horizontal direction, the clamping force N required for orderly conveying can be calculated using Equation (16).

2.1.4. Transmission Mechanism Design

The transmission mechanism, as was shown in Figure 16, was driven by a DC motor. The entire transmission system was divided into two groups: one for the bottom conveyor belt and the other for the vertical conveyor belts. During operation, Motor 1 drove the driving sprocket 2, which transmitted power via a chain to driven sprockets 3 and 4. Driven sprocket 4 rotated the horizontal drive roller 12, thereby driving the bottom conveyor belt to rotate—this constituted the first group of drive mechanisms. Simultaneously, driven sprocket 3 transmitted power through another chain to driven sprocket 5, which drove bevel gear set 6 to redirect the power, enabling synchronous pulley I to rotate at the same speed. Then, through timing belt 10, synchronous pulleys II and III rotated synchronously but in opposite directions, thus realizing the orderly vertical conveying of vegetables.

During field operations, the vegetable conveying capacity of the conveyor belts must be greater than or equal to the cutting output from the cutter. This requirement can be expressed as:

$$V_{\tau }={\displaystyle \frac{V_e{b}_1{q}_1}{h{b}_2{q}_2}}$$

(17)

where V_τ—conveying speed of the conveyor belt, m/s

V_e—forward speed of the machine, m/s

b₁—cutter working width, m

b₂—conveyor belt width, m

q₁—plant density in the field, plants/m²

q₂—plant density on the conveyor belt, plants/m²

h—stacking height of vegetables on the conveyor, m

Based on Equation (17), the required conveying speed V_τ falls within the range of 0.27 m/s to 0.65 m/s. Given that the diameter of the horizontal drive roller is 75 mm, the corresponding speed range of the drive roller can be calculated using Equation (18):

$$n={\displaystyle \frac{V_{\tau }}{2\pi r}}$$

(18)

where n—rotational speed of the horizontal drive roller, r/s

r—radius of the drive roller, m

Chain transmission is selected as the main driving method due to its compact structure, high reliability, and no slippage. Considering stability and size constraints, the number of teeth z₂ of driving sprocket 2 is set to 13, with a module m₂ of 4, addendum circle diameter d_a of 59 mm, and dedendum circle diameter d_f of 41 mm. The center distance a₁ between the driving sprocket and the driven sprocket 4 is 130 mm. The structural parameters of driven sprocket 4 are identical to those of the driving sprocket, and the transmission ratio is calculated by:

$$i_{24}={\displaystyle \frac{n_2}{n_4}}={\displaystyle \frac{Z_4}{Z_2}}$$

(19)

Since z₂ = z₁, the driving sprocket 2 and driven sprocket 4 rotate at the same speed. To ensure the same linear velocity for the bottom and vertical conveyor belts, and knowing that the diameter of the horizontal drive roller is 75 mm and the vertical drive roller is 65 mm, it can be known from Equation (19) that the number of teeth z₃ of driven sprocket 3 is calculated as 15, with an addendum diameter dₐ of 66 mm and center distance a₂ to the driving sprocket of 177 mm. According to the national standard GB/T 1243-2006, the chain pitch p is 12.7 mm, and the number of chain links is chosen to be even.

To achieve power transmission between intersecting axes, a straight bevel gear set is adopted for its simple structure and suitability for low-speed transmissions. Based on national standards GB/T 12368-1990 and GB/T 12369-1990, the bevel gear is designed with: number of teeth z₆ = 16; pressure angle α = 20°; addendum coefficient h_a* = 1; clearance coefficient c* = 0.2; and transmission ratio i = 1.

2.2. Test Method

Based on the structural and parametric determination of key components, a self-propelled orderly harvester for high-stem Chrysanthemum coronarium was developed. An orthogonal experimental design was used to conduct field trials to investigate the effects of operational parameters on harvesting performance and optimize parameter combinations.

2.2.1. Experimental Conditions

Field experiments were conducted using the developed prototype in a greenhouse facility at Yijiakang Ecological Agriculture Co., Ltd., Wuxi, China. The temperature that day was between 25 and 28 degrees Celsius, and the humidity ranged from 30% to 60%. The tested variety was Small-leaf Glebionis coronarium, supplied by Hefei Hefeng Seed Industry Co., Ltd., Hefei, China. A wide-ridge multi-row planting pattern was adopted under protected cultivation. The soil was clay loam. The ridge dimensions were as follows: ridge top width ≈ 900 mm, ridge base width ≈ 1000 mm, furrow width 200 mm, ridge height 200 mm. The ridge surface had ≥95% soil pulverization, was evenly leveled and compacted, and had a straightness consistency ≥95%.

Table 2. Test conditions.

Parameter	Value
Crop average height/mm	39.4
stem average diameter/mm	4.95
plot size/m2	500
soil moisture content/%	15~20
soil compactibility/kpa	3000–3500
Temperature/°C	25~28
Humidity/%	30~60

below is a summary table of all the field test conditions.

2.2.2. Experimental Method

The field surface conditions, soil conditions, and test site selection were based on the general provisions of GB/T 5262-2008 Method for Determination of Agricultural Machinery Test Conditions and DG/T 249-2024 Leafy Vegetable Harvester standards [40,41]. A row of high-stem Chrysanthemum coronarium cultivated under a standardized agronomic planting model was selected as the test site. A measuring tape was used to mark out a 35-meter-long test area, which was sequentially divided into three zones: a stabilization zone, a test zone, and a stopping zone. Within the test zone, three sampling areas of 1.8 m² (0.9 m × 2 m) were selected at 3 m intervals along the forward direction of the harvester. During the orthogonal performance test of the harvester, the number of high-stem Chrysanthemum coronarium plants in each sampling area was counted and recorded as N_i (i = 1, 2, 3).

According to the theoretical calculations of the harvester parameters for high-stem Chrysanthemum coronarium and the analysis of factors affecting harvesting quality, the primary parameters influencing the harvesting loss rate and orderliness include the cutting speed of the reciprocating cutter, the forward speed of the machine, and the conveying speed of the conveying mechanism. Therefore, in this experiment, three key operating parameters of the harvester were selected as influencing factors to conduct harvesting performance tests: machine forward speed (x₁), reciprocating cutting speed (x₂), and conveying speed of the conveying mechanism (x₃). The goal was to determine the optimal combination of operating parameters for the harvester.

Based on the actual operating speed range of the high-stem Chrysanthemum coronarium harvester (200–500 mm/s), the forward speed was set at three levels: 250 mm/s, 350 mm/s, and 450 mm/s. The reciprocating cutting speed ranged from 150 to 300 mm/s and was set at 150 mm/s, 200 mm/s, and 250 mm/s. The conveying speed of the conveying mechanism ranged from 270 to 650 mm/s and was set at 300 mm/s, 400 mm/s, and 500 mm/s. The orthogonal test design is shown in

Table 3. Factors and levels of orthogonal experiment.

Level	Factor
	x1/mm·s−1	x2/mm·s−1	x3/mm·s−1
1	250	150	300
2	350	200	400
3	450	250	500

.

After harvesting, performance indicators were calculated as follows:

(1)

Harvest Loss Rate

After the operation, damaged plants in the collection box and uncut, missed, or dropped plants in the test area were weighed. Harvest loss rate was calculated using:

$$S_l={\displaystyle \frac{M_l}{M}}\times 100\%$$

(20)

where S_l—harvest loss rate, %

M_l—weight of lost/damaged plants, kg

M—total weight of tall crown daisy in the test area, kg

(2)

Harvest Orderliness

The number of plants maintaining an upright, orderly posture from cutting to the delivery platform was recorded. Orderliness was calculated as:

$$S_s={\displaystyle \frac{N_s}{N}}\times 100\%$$

(21)

where Ss—harvest orderliness, %

Ns—number of upright, well-positioned plants

N—total number of harvested plants

3. Results

3.1. Experimental Results and Analysis

The machine’s forward speed, the reciprocating cutter’s cutting speed, and the conveyor speed were denoted as x₁, x₂, and x₃, respectively. The harvest loss rate and the harvest orderliness were represented as Y₁ and Y₂. A total of 17 test groups were conducted. The experimental design and results are shown in

Table 4. Response surface analysis results.

Test Number	Test Factors			Test Index
	x1/mm·s−1	x2/mm·s−1	x3/mm·s−1	Y1/%	Y2/%
1	450	250	400	8.01	63.24
2	250	200	500	5.51	80.15
3	350	150	500	6.94	86.93
4	250	200	300	5.69	86.9
5	350	150	300	5.32	81.86
6	450	150	400	9.28	70.13
7	350	250	500	4.65	67.12
8	350	200	400	9.85	87.91
9	350	250	300	5.29	88.28
10	350	200	400	10.23	89.49
11	450	200	500	8.28	70.57
12	250	150	400	6.35	77.03
13	350	200	400	10.16	87.86
14	450	200	300	8.18	71.1
15	350	200	400	10.13	86.61
16	250	250	400	6.46	81.19
17	350	200	400	10.21	85.62

Note: x₁—forward speed; x₂—cutting speed; x₃—conveyor speed; Y₁—harvest loss rate; Y₂—harvest orderliness.

.

The ANOVA results of the harvest loss rate are shown in

Table 5. Analysis of orthogonal test results of harvest loss rate.

Soruce of Variation	Quadratic Sum	Degree of Freedom	Mean Square Error	F Value	p Value
Model	65.50	9	7.27	106.26	<0.0001 **
x1	11.85	1	11.85	173.14	<0.0001 **
x2	1.51	1	1.51	22.10	0.0022 **
x3	0.10	1	0.10	1.47	0.2634
x1x2	0.47	1	0.47	6.95	0.0335 *
x1x3	0.02	1	0.02	0.28	0.6092
x2x3	1.27	1	1.27	18.64	0.0034 **
x12	1.58	1	1.58	23.10	0.0019 **
x22	16.47	1	16.47	240.53	<0.0001 **
x32	28.20	1	28.20	411.76	<0.0001 **
Residual error	0.48	7	0.07
Lack of fit	0.38	3	0.13	5.42	0.07
Pure error	0.09	4	0.02
Total	65.98	16

Note: “**” indicates extremely significant (p < 0.01); “*” indicates significant (p < 0.05).

. As observed, the regression model for harvest loss rate Y₁ was highly significant (p < 0.01), with a coefficient of determination R² = 0.90, which indicated that the model has high fitting accuracy and can reliably reflect the influence of various factors on harvest loss rate within the selected range. Based on the p-values, forward speed x₁ and reciprocating cutter speed x₂ had extremely significant effects on the model (p < 0.01), while conveyor speed x₃ had no significant effect. The regression equation relating the harvest loss rate to the coded variables is as follows:

$$\begin{array}{l}{Y}_1=10.12+1.22x_1-0.44x_2+0.11x_3-0.35x_1{x}_2+0.77x_1{x}_3\\ -0.57x_2{x}_3-0.61x_1^2-1.98x_2^2-2.59x_3^2\end{array}$$

(22)

According to the absolute values of the coefficients in Equation (22), the influence of each factor on the response variable can be ranked. The larger the absolute value, the more significant the influence. The absolute values of the coefficients for x₁, x₂, and x₃ are 1.22, 0.44, and 0.11, respectively, indicating the order of influence as forward speed x₁ > cutting speed x₂ > conveyor speed x₃.

Among the interaction terms, the interactions between x₁x₂ and x₂x₃ were significant, while the interaction between x₁x₃ was not significant (p = 0.6092). The response surface methodology (RSM) was used to explore the interactions among forward speed, cutting speed, and conveyor speed and their effects on the harvest loss rate. The corresponding 3D response surface plots are shown in Figure 2 and Figure 3.

As shown in Figure 17, the 3D response surface showed a steep slope, and the 2D contour plot presented an elliptical shape, indicating that the interaction between forward speed x₁ and cutting speed x₂ significantly affected the harvest loss rate, which was consistent with the ANOVA results. At different cutting speeds, the harvest loss rate increased with forward speed. Conversely, at different forward speeds, the harvest loss rate initially increased and then decreased as cutting speed changes, but this trend was more gradual. The denser contour lines suggested that forward speed had a more significant effect on harvest loss rate.

As shown in Figure 18, under different cutting speeds, the harvest loss rate increased first and then decreased as the conveyor speed increased. A similar trend was observed across different forward speeds. The density of the contour lines indicated that cutting speed has a more pronounced impact on harvest loss rate.

ANOVA was also conducted on the harvest orderliness data in Table 4, and the results are shown in

Table 6. Analysis of orthogonal experiment results of harvest orderliness.

Soruce of Variation	Quadratic Sum	Degree of Freedom	Mean Square Error	F Value	p Value
Model	1147.18	9	127.46	27.43	0.0001 **
x1	315.38	1	315.38	67.87	<0.0001 **
x2	32.48	1	32.48	6.99	0.0332 *
x3	68.27	1	68.27	14.69	0.0064 **
x1x2	30.53	1	30.53	6.57	0.1923
x1x3	9.67	1	9.67	2.08	0.0005 **
x2x3	172.00	1	172.00	37.02	0.0374 *
x12	359.02	1	359.02	77.27	<0.0001 **
x22	121.26	1	121.26	26.10	0.0014 **
x32	4.95	1	4.95	1.06	0.3365
residual error	32.53	7	4.65
Lack of Fit	23.94	3	7.98	3.72	0.1185
pure error	8.58	4	2.15
total	1179.71	16

Note: “**” indicates extremely significant (p < 0.01); “*” indicates significant (p < 0.05).

. As shown, the model’s significance level (p < 0.01) indicated the regression model is highly significant. The lack-of-fit test is not significant (p > 0.05), implying high model accuracy. The coefficient of determination R² = 0.93 indicates that 93% of the variation in harvest orderliness can be explained by the model, suggesting strong reliability.

Based on p-values, the forward speed x₁ and conveyor speed x₃ had an extremely significant impact on the model (p < 0.01), while cutting speed x₂ had a significant effect (p < 0.05). The regression equation relating the harvest orderliness to the coded variables is as follows:

$$\begin{array}{l}{Y}_2=87.50-6.28x_1-2.01x_2-2.92x_3-2.76x_1{x}_2+1.55x_1{x}_3\\ -6.56x_2{x}_3-9.23x_1^2-5.37x_2^2-1.08x_3^2\end{array}$$

(23)

From the ANOVA results, each factor’s effect varies. According to the principle that smaller p-values indicate more significant influence, the ranking is: forward speed x₁ > conveyor speed x₃ > cutting speed x₂. The interaction terms x₁x₃ and x₂x₃ are significant, while x₁x₂ is not.

Among the interaction terms, the interactions between x₁x₃ and x₂x₃ were significant, while the interaction between x₁x₂ was not significant. Response surface methodology was used to investigate the interactions among forward speed, cutting speed, and conveying speed, as well as their effects on harvesting orderliness. The corresponding three-dimensional response surface plots are shown in Figure 19 and Figure 20.

As shown in Figure 19, the elliptical 2D contour plot indicated a significant interaction between x₁ and x₂ on harvest orderliness, which aligned with the ANOVA results. When forward speed was constant, harvest orderliness first increased and then decreased with increasing cutting speed. When cutting speed was fixed, harvest orderliness exhibited a similar trend with forward speed. Contour line density showed that forward speed had a greater influence than cutting speed.

As shown in Figure 20, the interaction between forward speed x₁ and conveyor speed x₃ significantly affected harvest orderliness. When conveyor speed remains constant, harvest orderliness first increased and then decreased with increasing cutting speed. When cutting speed was fixed, harvest orderliness decreased as conveyor speed increased. The density of contour lines indicated that conveyor speed had a more pronounced effect on harvest orderliness.

3.2. Parameter Optimization

To optimize the operating parameters of the reciprocating cutting system, the harvesting loss rate (Y₁) and harvesting orderliness (Y₂) were used as objective functions. The Optimization module in Design-Expert 13 was utilized to optimize the model, with the objective set to minimize the harvesting loss rate and maximize the harvesting orderliness. The optimal parameter combination for harvesting high-stem Chrysanthemum coronarium was determined as follows: forward speed of 257.9 mm/s, cutting speed of 248.4 mm/s, and conveying speed of 300.45 mm/s. Under these conditions, the predicted harvesting loss rate and harvesting orderliness were 4.31% and 90.23%, respectively.

3.3. Field Validation Test

To facilitate field validation, the optimized parameters were appropriately rounded: the forward speed was set to 260 mm/s, the cutting speed to 250 mm/s, and the conveying speed to 300 mm/s. Field experiments were conducted based on these rounded values. Each test was repeated three times, and the average values were calculated. The operational performance of the machine in the field is shown in Figure 21.

The test results showed that the overall performance of the prototype was relatively stable, and all mechanisms functioned properly. The damage and loss to high-stem Chrysanthemum coronarium plants were minimal, and the harvester demonstrated a high level of orderliness in harvesting, generally meeting the mechanized harvesting requirements of row-sown high-stem Chrysanthemum coronarium. According to the aforementioned testing method, the harvesting loss rate and orderliness of each test area were calculated, and the average of three trials was taken. The detailed field experiment results are shown in

Table 7. Results of field experiment under standardized agronomic planting pattern.

Test Number	Harvesting Loss/%	Harvesting Orderliness/%
1	4.56	91.78
2	4.47	92.47
3	4.32	93.46
Average value of the test	4.45	92.57
Simulation prediction value	4.31	90.23
relative error	3.25	2.59

. The average harvesting loss rate of the prototype was 4.45%, and the average harvesting orderliness was 92.57%. The relative error compared to the predicted values was less than 5%, which meets the design requirements.

4. Discussion

Through the structural design of key components, optimization of working parameters, and validation via field trials, the self-propelled high-stem Chrysanthemum coronarium orderly harvester developed in this study performed well in critical operations such as row separation, cutting, orderly conveying, and collection. It successfully achieved the goal of low-damage, flexible, and orderly harvesting. In terms of stem cutting, based on cutting mechanics analysis, kinematic analysis, and numerical simulation, the designed cutting device ensured stability and low damage during cutting. In the stem conveying stage, the kinematics and dynamics of pre-cut clamping, clamping-cutting, and orderly conveying stages were thoroughly analyzed, and the optimized conveying mechanism effectively maintained the orderly posture of the plants. For the orderly collection stage, this study adopted a combination of a deflector plate and a floating collection basket, replacing the free-fall method used by most existing leafy vegetable harvesters. Field test data clearly showed that this approach significantly reduced plant posture inversion and disorder during collection, thereby improving the final degree of orderliness inside the collection basket and laying a solid foundation for subsequent automated handling.

Compared with the vegetable harvesters developed by Yue Jin and Dengyu Xiong, the orderly cutting and harvesting method designed in this study also achieved efficient harvesting of high-stem Chrysanthemum coronarium [42,43]. It showed certain advantages in terms of reduced harvesting damage and improved orderliness. These benefits were mainly attributed to the agronomic considerations for high-stem Chrysanthemum coronarium in field cultivation, as well as the theoretical analysis conducted on the reciprocating cutting and orderly conveying mechanisms, which effectively enhanced the operational performance of the machine.

The mechanical analysis of stem cutting, kinematic analysis of the mechanism, numerical simulation, the geometric and kinematic conditions along the conveying path, and the mechanical model of the conveying process for high-stem Chrysanthemum coronarium were comprehensively studied. The key structural and working parameters of the harvester were determined, which constitute the innovations of this research.

However, this study also had certain limitations and identified directions for further exploration, thus pointing the way for future research:

Fine optimization of core components: Although the current design met the basic harvesting requirements for high-stem Chrysanthemum coronarium, there is still room for improvement in the cutting device (e.g., blade gap and angle matching, adaptability to different stem hardness) and in the conveying process (e.g., stem slippage at high speeds, entanglement control, and anti-blocking design for denser or more tender stems). Future work should focus on enhancing cutting consistency, reducing power consumption, and improving the stability and low-damage performance of the conveying system to adapt to a wider range of field conditions, such as adding automatic row guidance or a clamping system that can be adjusted for different plant sizes.

Expansion and validation of general applicability: The current field experiments mainly focused on high-stem Chrysanthemum coronarium. To enhance the general value and application potential of the technology, it is necessary to extend the research to other representative leafy vegetables such as bok choy and spinach. These crops differ significantly in plant morphology (plant height, leaf expansion), stem mechanical properties (e.g., fiber content, toughness), and harvesting requirements. Future studies should analyze how these physical differences affect key mechanisms (e.g., cutting height adjustment range, clamping/limiting devices, and deflector structures), and use modular design or parameter-adjustable mechanisms to verify and improve the adaptability and operational efficiency of the harvester across different vegetable varieties.

Integration of intelligent perception and adaptive control: Parameter optimization in this study was based on preset field conditions. The next step will involve using visual sensing technology to perceive crop density, height, and growth posture in real time, which can provide accurate online adjustments for key operational parameters such as cutting height, cutting frequency and force, and conveyor belt speed and inclination. This will enable adaptive control of the system, thereby improving the consistency of harvesting quality and adaptability to fields with varying growth conditions.

In conclusion, this study successfully validated the feasibility of the self-propelled high-stem Chrysanthemum coronarium orderly harvester and demonstrated the advantages of its core innovations. Future optimization should focus on enhancing the performance of core mechanisms, expanding multi-variety applicability, integrating intelligent perception and control strategies, and ultimately achieving fully unmanned harvesting operations. This is not only a necessary step for upgrading the equipment but also an important contribution to the advancement of efficient, low-damage mechanized harvesting technology for leafy vegetables. The prototype trial production cost is approximately 80,000 yuan. By adopting mechanized harvesting, each mu can save about 150 yuan in costs. If 200 mu of tall-stemmed sorghum are harvested twice a year, it can recover the cost within 2 years. With the prototype being finalized and mass production underway, the manufacturing cost can be further reduced.

5. Conclusions

(1)

Based on the agronomic requirements for high-stem Chrysanthemum coronarium planting and the mechanized orderly harvesting requirements, a self-propelled high-stem Chrysanthemum coronarium orderly harvester was developed. The machine mainly consists of an electric four-wheel chassis, a row divider, a reciprocating cutting mechanism, an orderly conveying mechanism, a deflector plate, and a floating collection platform. It was capable of completing multiple harvesting operations—such as row separation, cutting, orderly conveying, and collection—in a one-pass operation.

(2)

A reciprocating electric cutting mechanism based on an eccentric cam was designed, and its kinematic model was established. The eccentricity of the cam was determined to be 8.5 mm, resulting in a cutting stroke of 17 mm and a crank length of 57 mm. Under the constraint that the dual moving blades must clamp and cut the stem without slipping, the sliding-cut angle was calculated as 19°, with the blade width C = 14 mm, d = 7 mm, and blade height h = 20 mm. Using ADAMS software, a kinematic simulation of the cutting mechanism under no-load conditions was conducted. The displacement, velocity, and acceleration curves of the two blades were identical in magnitude but differed in phase by π, allowing their inertial forces to cancel each other out and thereby reducing vibrations transmitted to the cutting table.

(3)

The geometric and kinematic conditions for orderly conveying of high-stem Chrysanthemum coronarium along the transport path were analyzed, and a mechanical model of the conveying process was established. The key operating parameters affecting conveying performance were identified: the clamping and conveying angle ranged from 20° to 30°, and the conveying speed ranged from 0.27 to 0.65 m/s. The transmission system was theoretically designed, with key structural parameters determined to ensure equal linear velocity between the lower conveyor belt and the transverse conveyor belt.

(4)

Field tests of the orderly harvesting of high-stem Chrysanthemum coronarium were conducted. Regression models were established with harvesting loss rate and orderliness as evaluation indicators. Through variance analysis and regression analysis, the effects of cutting speed, forward speed of the machine, conveying speed, and their interaction on the evaluation indicators were revealed.

(5)

A regression model was established for harvesting loss rate and harvesting orderliness, followed by optimization. The optimal parameter combination was determined as follows: forward speed of 260 mm/s, cutting speed of 250 mm/s, and conveying speed of 300 mm/s. Validation tests showed that the average harvesting loss rate of the harvester prototype was 4.45%, and the average orderliness was 92.57%. The relative error compared to the predicted values was less than 5%, meeting the orderly harvesting requirements of high-stem Chrysanthemum coronarium. Compared with manual harvesting, mechanized harvesting can increase the harvesting efficiency by more than 10 times, and can save approximately 150 yuan per mu of cost.