3D Modeling Simulation and Innovative Design of a Saw Cutting Mechanism for a Movable Buckling Machine

Abstract

In this research, a sawing mechanism of a mobile buckling machine was innovatively designed for the forest resources and environmental characteristics of the northeast region of China. The northeast region of China is rich in forest resources, but the climate is cold, which puts high demands on the performance and adaptability of mechanical equipment. In this study, the three-dimensional modeling of the saw-cutting mechanism was completed by Pro/Engineer 5.0 modeling software, and an in-depth dynamics simulation analysis was carried out using dynamics simulation software. The study aims to assess the applicability of this saw-cutting mechanism in the northeast region of China, analyze its design characteristics, and explore possible directions for performance optimization. Through this study, we expect to provide valuable references for future technological innovations, as well as to promote the practical application and development of mobile buckling machines in cold regions. This research not only has regional characteristics, but also reflects the innovative idea of combining mechanical equipment design with environmental adaptability.

1. Introduction

The Northeast region of China is known as the “Green Treasure Trove” for its vast forest cover and abundant forest resources. This fertile land in China’s high latitude region not only nurtures diverse biological populations, but also provides a steady stream of high-quality raw materials for China’s and the world’s timber markets. However, as the global warming trend intensifies, the cold climate characteristics of the Northeast region are becoming more and more prominent, bringing unprecedented challenges to the sustainable utilization of local forest resources.

Traditional buckling machines play an important role in the development of forest resources, but their performance stability and work efficiency under extreme climatic conditions are greatly limited [1]. In the severe winters, traditional buckling machines face many problems such as the easy freezing of mechanical parts, the increased viscosity of lubricating oil, and difficulty in starting the equipment, which not only affects the progress of forest harvesting, but also increases the operating costs and safety hazards. Therefore, the development of a mobile buckling machine sawing mechanism that can adapt to the high-latitude cold environment, with high stability and reliability, has become an urgent need for the development of current forestry machinery technology.

A new type of forestry machinery and equipment is the mobile buckling machine sawing mechanism, the design concept of which is aimed at solving the limitations of traditional buckling machines that are working in harsh environments [2,3,4]. Through the introduction of advanced material science, mechanical design, and intelligent control technology, the mobile buckling machine saw-cutting mechanism can still maintain efficient operation in low-temperature conditions, which not only improves the efficiency and safety of forest logging, but also reduces the degree of damage to the environment [5]. In addition, its strong mobility also means the equipment can flexibly adapt to different terrains and operational needs, providing a strong guarantee for the rational use of forest resources.

In the context of environmental protection and sustainable development, the importance of the mobile buckling machine saw-cutting mechanism is self-evident. It not only helps to realize the sustainable utilization of forest resources, but also can protect the integrity and diversity of ecosystems while promoting local economic development. Therefore, this study will deeply explore the application prospects and key technologies of the mobile buckling machine sawing mechanism in the utilization of forest resources in the northeast region, with a view to providing new ideas and methods for the development of forestry machinery technology [6].

2. Materials and Methods

2.1. Materials

When contrasting the superiority of Q345 as a material for sawing mechanisms, we have considered two common metallic materials as points of reference: A36 steel and aluminum alloy (such as 6061-T6 aluminum alloy), with a particular focus on their performance in cold climates to further substantiate the suitability of Q345.

2.1.1. A36 Steel

A36 steel, a commonly used structural steel, exhibits favorable weldability, processability, and a relatively low cost. Nevertheless, under cold climatic conditions, its mechanical properties fall short of those of Q345. Specifically:

• Strength Differential: While the yield and tensile strengths of A36 steel suffice for general structural requirements, they are comparatively lower than those of Q345. Under extreme cold conditions, the material’s properties may further deteriorate, leading to inadequate stability of the sawing mechanism under heavy loads and posing potential safety hazards.
• Toughness Issue: A36 steel exhibits poor impact toughness at low temperatures, rendering it more susceptible to brittle fracture. This is a non-negligible drawback for sawing mechanisms that frequently operate in cold environments. In contrast, Q345 maintains good toughness even at low temperatures, effectively preventing the occurrence of brittle fracture.
• Corrosion Resistance: A36 steel possesses relatively weak corrosion resistance, making it vulnerable to corrosion in wet or salty cold environments, which can subsequently shorten its service life. In contrast, Q345, with its superior corrosion resistance, is better suited to withstand harsh working conditions.

2.1.2. 6061-T6 Aluminum Alloy

Aluminum alloys are widely utilized across various industries due to their lightweight nature, high strength, and excellent corrosion resistance. However, in the context of sawing mechanisms, particularly those subjected to significant stress and operating in cold climates, aluminum alloys are not the optimal choice:

4.

Strength and Rigidity: While aluminum alloys exhibit a high specific strength (ratio of strength to weight), their absolute strength, such as yield strength and tensile strength, is generally lower than that of Q345. During sawing processes, particularly when dealing with hard materials or requiring high stability, aluminum alloys may not meet the necessary requirements.

5.

Low-Temperature Performance: Although aluminum alloys do not become extremely brittle at low temperatures like some metals, they still exhibit a notable decrease in strength as the temperature drops. In contrast, Q345 maintains relatively high strength and toughness even at low temperatures, making it more suitable for operational demands in cold climates.

6.

Wear Resistance: Sawing mechanisms frequently come into contact with materials during operation, leading to wear and tear. Aluminum alloys possess relatively poor wear resistance, which may result in faster wear and more frequent maintenance requirements. In contrast, Q345 exhibits better wear resistance, contributing to the extended service life of sawing mechanisms.

In summary, Q345, as a material for sawing mechanism, exhibits notable advantages. Its superior mechanical properties, including high yield strength and tensile strength, ensure the stability and safety of the saw-cutting mechanism under considerable stress. Moreover, Q345 retains and adapts to various working conditions. Additionally, its excellent corrosion resistance and processability enable Q345 to withstand harsh working environments and be easily fabricated into various shapes and structures of saw-cutting mechanism components, thereby, enhancing the overall structural stability and reliability. These advantages render Q345 an ideal material for the manufacturing of high-performance saw-cutting mechanisms.

The definition of material properties is presented in

Table 1. Material Properties of the Saw Cutting Mechanism.

Material	Material Density (kg/m3)	Elastic Modulus (Pa)	Poisson’s Ratio	Yield Strength (MPa)
Q345	7.85 × 103	2.1 × 1011	0.3	345

.

2.2. Methods

2.2.1. Structural Design of the Saw-Cutting Mechanism

The saw-cutting mechanism serves as the primary execution component of the buckling machinery, primarily tasked with the precision sawing of logs during the buckling process. Our objective is to design a wood buckling machine that can operate effectively in low-temperature environments, exhibiting flexibility, high intelligence, efficiency, and minimal resource consumption. The main structure of this machine is illustrated in Figure 1.

When the logs are positioned for sawing, the wood is securely clamped in place by the clamping device, actuated precisely by the hydraulic system. As the sawing device advances along the guide rail, it gradually feeds the wood into the saw-cutting mechanism, effectively cutting it off [7,8]. Upon completion of the sawing process, the clamping device releases its grip on the log, once again controlled precisely by the hydraulic system. The severed portion of the log is then transported to the designated location for further processing [9]. Finally, the sawing device retracts back to its initial position along the guide rail, guided accurately by the hydraulic system, completing a cycle of the sawing operation.

The sawing device primarily comprises a motor, a reducer, a fixing element, and a wood-sawing mechanism. The motor, via the reducer, drives the rotation of the sprocket wheel. The wood-sawing mechanism is predominantly constituted by a sprocket wheel, a guide, and chain saws [10]. The structural configuration of this device is depicted in Figure 2.

2.2.2. Three-Dimensional Solid Model Building

By delving into the construction of the solid model for the saw-cutting mechanism, we have thoroughly examined the intricate processes of component modeling and overall assembly [11]. This process is facilitated by the advanced Pro/Engineer 5.0 (formerly known as Pro/Engineer, now renamed Creo Parametric) software platform, which leverages its parametric design and solid modeling techniques. Through these cutting-edge technologies, we have precisely modeled each individual component of the saw-cutting mechanism and conducted a meticulous overall assembly, ultimately yielding a high-precision three-dimensional model as depicted in Figure 3 [12]. This model not only accurately reflects the physical structure of the saw-cutting mechanism, but also provides a solid digital foundation for subsequent simulation analysis, optimization design, and practical manufacturing.

During the assembly process in Pro/E, we employed a dynamic assembly approach. Initially, we precisely defined the critical moving components to ensure that their kinematic properties align with the requirements for subsequent dynamic simulations [13]. Upon completion of the assembly, we conducted rigorous interference detection on the model, aiming to comprehensively examine the presence of potential interferences or conflicts within the model [14]. In the event of detected interference, we promptly corrected the involved components and reassembled them, iterating this process until all interferences were completely eliminated, thereby, ensuring the accuracy and reliability of the model.

2.2.3. Static Analysis

In this study, we employed the Pro/Engineer 5.0 software to generate a three-dimensional model of the workbench and successfully imported the fixed components of the saw and hydraulic cylinder from the Pro/Engineer environment into Pro/Mechanica 5.0 for indepth mechanical analysis [15,16]. To accurately evaluate the strength performance of these components, we specifically selected equivalent stress ($\mathrm{\sigma }r$) as the evaluation metric. This metric comprehensively reflects the overall stress level of the components under complex stress states, thereby, providing us with reliable data regarding their strength performance. Through this approach, we aim to optimize product design and ensure its stability and safety under various working conditions.

The yield strength of Q345 steel is specified as 345 MPa, and the permissible stress is as follows:

$$[\sigma ]=\sigma /\mathrm{S}=345/1.5=230(\mathrm{MPa})$$

(1)

The grid division of the crucial components, as depicted in Figure 4, is carried out with precision.

After a thorough analysis of the material properties and strength requirements, we employ the equivalent stress intensity evaluation method for precise assessment. If the equivalent stress $\mathrm{\sigma }\mathrm{r}$ does not exceed the permissible stress $[\mathrm{\sigma }]$ of the material, it can be concluded that the structural strength satisfies the predefined requirements.

2.2.4. Saw-Cutting Mechanism Simulation Analysis

Firstly, the three-dimensional model was precisely imported into the “agencies” analysis module of Pro/E to ensure the accuracy and integrity of the data [17,18]. Subsequently, we rigorously defined various physical performance parameters of the model’s material to guarantee the precision of the simulation. Then, based on the actual application scenario, we accurately set the performance parameters of the servo motor. Furthermore, we needed to define the potential forces and torques that the model may encounter during motion, in order to simulate the mechanical environment under real-world conditions. Finally, by incorporating dynamic parameters such as damping, we enhanced the description of the model’s dynamic characteristics, thus, providing a comprehensive and accurate preparation for subsequent simulation analysis [19].

3. Results and Discussion

3.1. Design Derivation for the Saw-Cutting Mechanism

When birch serves as the subject of sawing, the cutting speed $v$ can be precisely calculated.

$$v=\frac{nzt}{3000i}$$

(2)

In the formulation, where the variable $v$ represents the cutting speed in meters per second (m/s), n denotes the motor speed in revolutions per minute (r/min), z signifies the number of teeth on the drive sprocket, t represents the pitch of the sprocket in millimeters (mm), and i is the gear transmission ratio, the calculated cutting speed of the cutting device is precisely 5.5 mm per second (mm/s).

Feed speed $v_H$:

$${\displaystyle \frac{t^{\prime }}{v}}={\displaystyle \frac{m}{v_H}}$$

(3)

$$v_H={\displaystyle \frac{vm}{t^{\prime }}}={\displaystyle \frac{5.5\times 0.0002}{0.01026\times 4}}=0.0268(\mathrm{m}/\mathrm{s})$$

(4)

In the formulation, v_H represents the feed speed expressed in meters per second (m/s), while $v$ signifies the cutting speed, also measured in meters per second (m/s). t′ designates the pitch of the sprocket, and m denotes the feed per tooth.

Cutting resistance F:

$$F={\displaystyle \frac{KLb{v}_H}{v}}(\mathrm{kg})$$

(5)

$$F={\displaystyle \frac{KLb{v}_H}{v}}={\displaystyle \frac{3.714\times 7.5\times 0.02}{5.5}}L=0.1013L(kg)=0.9927L(N)$$

(6)

In the formulation, the variable K represents the unit cutting resistance, expressed in kg·m/cm³. v_H denotes the feed speed, measured in meters per second (m/s). The cutting speed denoted by $v$, is also expressed in meters per second (m/s). L represents the kerf length, measured in millimeters (mm). b signifies the kerf width, also in millimeters (mm).

Feed force F_H:

$$F_H={\alpha }_{0}F=0.6\times 0.9927L=0.5956(L)$$

(7)

In the formulation, the variable ${\mathrm{\alpha }}_0$ represents the saw tooth-related coefficient. L represents the kerf length, measured in millimeters (mm).

The rolling guide section is depicted in Figure 5 and the feed mechanism consists of a workbench and guide rail [20,21,22].

Figure 6 illustrates the guide rail and slider in conjunction.

A detailed analysis of the cutting mechanism is presented in Figure 7 and Figure 8.

When the saw-cutting mechanism is in a non-cutting state, the radial load exerted on the guide rail is denoted as “P_urn”:

$$P_{ur1}={\displaystyle \frac{m_{1}g{l}_2}{2l_0}}+{\displaystyle \frac{m_{2}g{l}_3}{2l_0}}={\displaystyle \frac{42\times 9.8\times 0.036}{0.15\times 2}}+{\displaystyle \frac{19.5\times 9.8\times 0.03}{0.15\times 2}}=68.5(N)$$

(8)

$$\hspace{1em}\hspace{1em}{P}_{ur2}=-{\displaystyle \frac{m_{2}g{l}_3}{2l_0}}-{\displaystyle \frac{m_{1}g{l}_2}{2l_0}}=-{\displaystyle \frac{19.5\times 9.8\times 0.03}{0.15\times 2}}-{\displaystyle \frac{42\times 9.8\times 0.036}{0.15\times 2}}=-68.5(N)$$

(9)

$$\hspace{1em}\hspace{1em}{P}_{ur3}=-{\displaystyle \frac{m_{2}g{l}_3}{2l_0}}-{\displaystyle \frac{m_{1}g{l}_2}{2l_0}}=-{\displaystyle \frac{19.5\times 9.8\times 0.03}{0.15\times 2}}-{\displaystyle \frac{42\times 9.8\times 0.036}{0.15\times 2}}=-68.5(N)$$

(10)

$$P_{ur4}={\displaystyle \frac{m_{1}g{l}_2}{2l_0}}+{\displaystyle \frac{m_{2}g{l}_3}{2l_0}}={\displaystyle \frac{42\times 9.8\times 0.036}{0.15\times 2}}+{\displaystyle \frac{19.5\times 9.8\times 0.03}{0.15\times 2}}=68.5(N)$$

(11)

When the saw-cutting mechanism is in a non-cutting state, the lateral load exerted on the guide rail is referred to as “P_uhn”:

$$P_{uh1}={\displaystyle \frac{m_{1}g{l}_1}{2l_0}}={\displaystyle \frac{19.5\times 9.8\times 0.1}{2\times 0.15}}=63.7(\mathrm{N})$$

(12)

$$\hspace{1em}\hspace{1em}{P}_{uh2}=-{\displaystyle \frac{m_{1}g{l}_1}{2l_0}}=-{\displaystyle \frac{19.5\times 9.8\times 0.1}{2\times 0.15}}=-63.7(\mathrm{N})$$

(13)

$$\hspace{1em}\hspace{1em}{P}_{uh3}=-{\displaystyle \frac{m_{1}g{l}_1}{2l_0}}=-{\displaystyle \frac{19.5\times 9.8\times 0.1}{2\times 0.15}}=-63.7(\mathrm{N})$$

(14)

$$P_{uh4}={\displaystyle \frac{m_{1}g{l}_1}{2l_0}}={\displaystyle \frac{19.5\times 9.8\times 0.1}{2\times 0.15}}=63.7(\mathrm{N})$$

(15)

When the saw-cutting mechanism is in the cutting operation, the radial load exerted on the guide rail is denoted as ”P_urn”:

$$\begin{array}{l}{P}_{ur1}={\displaystyle \frac{m_{1}g{l}_2}{2l_0}}+{\displaystyle \frac{m_{2}g{l}_3}{2l_0}}+{\displaystyle \frac{F_H{l}_4}{2l_0}}+{\displaystyle \frac{F_l^5}{2l_0}}\\ ={\displaystyle \frac{42\times 9.8\times 0.036}{0.15\times 2}}+{\displaystyle \frac{19.5\times 9.8\times 0.03}{0.15\times 2}}+{\displaystyle \frac{400\times 0.4}{0.15\times 2}}+{\displaystyle \frac{700\times 0.096}{0.15\times 2}}\\ =825.8(N)\end{array}$$

(16)

$$\begin{array}{l}{P}_{ur2}=-{\displaystyle \frac{m_{1}g{l}_2}{2l_0}}-{\displaystyle \frac{m_{2}g{l}_3}{2l_0}}-{\displaystyle \frac{F_H{l}_4}{2l_0}}-{\displaystyle \frac{F_l^5}{2l_0}}\\ =-{\displaystyle \frac{42\times 9.8\times 0.036}{0.15\times 2}}-{\displaystyle \frac{19.5\times 9.8\times 0.03}{0.15\times 2}}-{\displaystyle \frac{400\times 0.4}{0.15\times 2}}-{\displaystyle \frac{700\times 0.096}{0.15\times 2}}\\ =-825.8(N)\end{array}$$

(17)

$$\begin{array}{l}{P}_{ur3}=-{\displaystyle \frac{m_{1}g{l}_2}{2l_0}}-{\displaystyle \frac{m_{2}g{l}_3}{2l_0}}-{\displaystyle \frac{F_H{l}_4}{2l_0}}-{\displaystyle \frac{F_l^5}{2l_0}}\\ =-{\displaystyle \frac{42\times 9.8\times 0.036}{0.15\times 2}}-{\displaystyle \frac{19.5\times 9.8\times 0.03}{0.15\times 2}}-{\displaystyle \frac{400\times 0.4}{0.15\times 2}}-{\displaystyle \frac{700\times 0.096}{0.15\times 2}}\\ =-825.8(N)\end{array}$$

(18)

$$\begin{array}{l}{P}_{ur4}={\displaystyle \frac{m_{1}g{l}_2}{2l_0}}+{\displaystyle \frac{m_{2}g{l}_3}{2l_0}}+{\displaystyle \frac{F_H{l}_4}{2l_0}}+{\displaystyle \frac{F_l^5}{2l_0}}\\ ={\displaystyle \frac{42\times 9.8\times 0.036}{0.15\times 2}}+{\displaystyle \frac{19.5\times 9.8\times 0.03}{0.15\times 2}}+{\displaystyle \frac{400\times 0.4}{0.15\times 2}}+{\displaystyle \frac{700\times 0.096}{0.15\times 2}}\\ =825.8(N)\end{array}$$

(19)

When the saw-cutting mechanism is in the cutting state, the lateral load exerted on the guide rail is denoted as “P_uhn”:

$$P_{uh1}={\displaystyle \frac{F{l}_6}{2l_0}}-{\displaystyle \frac{m_1{l}_1}{2l_0}}={\displaystyle \frac{400\times 0.15}{2\times 0.15}}-{\displaystyle \frac{19.5\times 9.8\times 0.1}{2\times 0.15}}=136.3(\mathrm{N})$$

(20)

$$\hspace{1em}\hspace{1em}{P}_{uh2}=-{\displaystyle \frac{F{l}_6}{2l_0}}+{\displaystyle \frac{m_1{l}_1}{2l_0}}=-{\displaystyle \frac{400\times 0.15}{2\times 0.15}}+{\displaystyle \frac{19.5\times 9.8\times 0.1}{2\times 0.15}}=-136.3(\mathrm{N})$$

(21)

$$\hspace{1em}\hspace{1em}{P}_{uh3}=-{\displaystyle \frac{F{l}_6}{2l_0}}+{\displaystyle \frac{m_1{l}_1}{2l_0}}=-{\displaystyle \frac{400\times 0.15}{2\times 0.15}}+{\displaystyle \frac{19.5\times 9.8\times 0.1}{2\times 0.15}}=-136.3(\mathrm{N})$$

(22)

$$P_{uh4}={\displaystyle \frac{F{l}_6}{2l_0}}-{\displaystyle \frac{m_1{l}_1}{2l_0}}={\displaystyle \frac{400\times 0.15}{2\times 0.15}}-{\displaystyle \frac{19.5\times 9.8\times 0.1}{2\times 0.15}}=136.3(\mathrm{N})$$

(23)

3.2. Static Analysis Results

The Stress Contour Map is presented in Figure 9, providing a visual representation of the stress distribution.

Through an in-depth analysis of the stress contour map of the hydraulic cylinder’s fixed component, we have precisely calculated its equivalent stress value to be 143.7 MPa (${\mathrm{\sigma }}_r=143.7\mathrm{MPa}$). Given the known upper limit of the permissible stress, which is 230 MPa ($[\mathrm{\sigma }]=230\mathrm{MPa}$), we have rigorously verified that the equivalent stress of the hydraulic cylinder’s fixed components (${\mathrm{\sigma }}_r$) is significantly below the threshold of the permissible stress $[\mathrm{\sigma }]$. This conclusion robustly demonstrates that the structural strength of the hydraulic cylinder meets the predetermined safety requirements, thereby ensuring its stability and reliability during operation.

Through a thorough analysis of the stress contour map of the workbench, we have precisely calculated its equivalent stress value at 104.7 MPa (${\mathrm{\sigma }}_r=104.7\mathrm{MPa}$). Given the established upper limit of permissible stress at 230 MPa ($[\mathrm{\sigma }]=230\mathrm{MPa}$), we are able to conclusively verify that the equivalent stress is significantly below the threshold of the permissible stress. This finding effectively demonstrates that the structural strength of the workbench not only meets but exceeds the predetermined safety requirements, thus, ensuring its stability and reliability during operation.

Through a rigorous analysis of the stress contour map of the saw’s fixed components, we have precisely determined its equivalent stress value to be 32.24 MPa (${\mathrm{\sigma }}_{\mathrm{r}}=32.24\mathrm{MPa}$). Given the established upper limit of the material’s permissible stress ($[\mathrm{\sigma }]=230\mathrm{MPa}$), which is 230 MPa, we have rigorously verified that the equivalent stress borne by the fixed components of the saw is significantly below the threshold of the permissible stress. This result forcefully demonstrates that the structural design of the saw not only fulfills the predetermined safety standards in terms of strength, but also possesses a notable degree of redundancy, thus, ensuring stability and reliability in practical working environments.

3.3. Saw-Cutting Mechanism’s Simulation Analysis Results

In the saw chain, a point P₁ is selected, and the velocity variations of P₁ in the X, Y, and Z-axis directions are depicted in Figure 10a–c. Furthermore, a point P₂ is identified on the guide plate. Since the velocity of P₂ remains constant in the X and Z-axis directions, the focus is on analyzing the velocity change of P₂ in the Y-axis direction. The velocity variations of point P₂ in the Y-axis direction are illustrated in Figure 10d.

As depicted in Figure 10a, the velocity of P₁ in the X-axis direction is null, indicating that the sawing device does not undergo positional changes in the X-axis direction. Based on Figure 10c, the cutting speed of the saw chain is determined to be 5.5 m/s. Additionally, Figure 10d reveals that the feed speed during the work process is 20 mm/s.

The hydraulic cylinder external axial load change is shown in Figure 11. Between 4 s and 30 s, the axial load of the hydraulic cylinder underwent variations. During this phase, the saw-cutting mechanism was in the cutting stage, during which significant changes in the axial load were observed.

4. Conclusions

In the application of sawing technology, the saw-cutting mechanism has ingeniously employed the buckling technique of chain saws, significantly reducing its volume and weight and greatly enhancing the portability of the equipment, and effectively minimizing wood waste during processing. Through a rigorous static analysis of the key components of the saw-cutting mechanism, we have determined that the equivalent stress is substantially lower than the permissible stress of the material, which conclusively validates the reliability and safety of the structural strength of the components. Additionally, we have conducted an in-depth dynamic analysis of the sawing mechanism, comprehensively grasping its kinematic characteristics and precisely mapping the dynamic variation curve of external axial loads during the sawing process. These research findings not only provide crucial insights for the optimization of the sawing mechanism but also significantly facilitate the design of hydraulic systems, contributing to the enhancement of the overall performance and efficiency of the sawing system.

The mobile buckling machine saw-cutting mechanism designed in this research has significant applicability and performance advantages. Through the comprehensive use of advanced modeling techniques and dynamics simulation analysis, this study successfully optimized the design of the sawing mechanism and verified its high efficiency, precision, and durability in the actual forest environment. The research results not only promote the application and development of mobile sawing machinery in cold regions, but also provide new directions and methods for the intelligent and environmentally adaptive design of future mechanical equipment.