Compressor and refrigeration apparatus

Abstract

The application discloses a compressor and refrigeration equipment, and relates to the technical field of compressors, wherein the compressor comprises a cylinder body, a roller, a crankshaft and a sliding vane, the cylinder body is internally provided with a working cavity and a sliding vane groove; the roller is eccentrically rotatable in the working cavity through the crankshaft, and the outer circumferential wall of the roller is concavely provided with an accommodating groove penetrating along the axial direction thereof; the sliding vane comprises a sliding vane main body and a sliding vane head, the sliding vane main body is reciprocable in the sliding vane groove, and the sliding vane head is hinged in the accommodating groove; the total length of the sliding vane in the radial direction is defined as Lv, the thickness of the sliding vane extending in the circumferential direction of the cylinder body is t, the outer diameter of the sliding vane head is D1, the eccentricity of the crankshaft is e, the cylinder height and the inner diameter of the cylinder body are H and D4 respectively, and the following conditions are met: 0.09<=e / t<=2.2; the technical scheme provided by the application can improve the reliability of the compressor.

Description

Technical Field

[0001] This invention relates to the field of compressor technology, and in particular to a compressor and refrigeration equipment. Background Technology

[0002] Roller compressors rely on eccentric rollers and the cylinder to form a changing working chamber to compress gas. Sliding vanes reciprocate within vane grooves to separate high and low pressure chambers and maintain a seal. However, existing vane designs suffer from inadequate key geometric parameters that fail to match the eccentricity and displacement. This makes the vanes highly susceptible to failure due to stress concentration, excessive wear, or fatigue, thus affecting the compressor's sealing performance, efficiency, and service life. Summary of the Invention

[0003] The main objective of this invention is to provide a compressor and refrigeration equipment that aims to improve the reliability of the compressor.

[0004] To achieve the above objectives, the compressor proposed in this invention comprises: The cylinder body has a working chamber inside, and the cylinder body also has a sliding vane groove that communicates with the working chamber; A roller and a crankshaft, wherein the roller is eccentrically rotatable within the working cavity via the crankshaft, and the outer peripheral wall of the roller is recessed with a receiving groove extending along its axial direction; A slider includes a slider body and a slider head connected to each other. The slider body can reciprocate within the slider groove, and the slider head is adapted to be installed and hinged within the receiving groove. Define the total radial length of the slide vane as Lv, the thickness of the slide vane extending circumferentially within the cylinder body as t, the outer diameter of the slide vane head as D1, the eccentricity of the crankshaft as e, and the cylinder height and inner diameter of the cylinder body as H and D4, respectively, satisfying: 0.09≤[10 -3 *e*H*π*((D4)^2-(D4-2e)^2)] / (4*Lv* )≤2.2.

[0005] In one embodiment, 0.37 ≤ [10 -3 *e*H*π*((D4)^2-(D4-2e)^2)] / (4*Lv* )≤1.05.

[0006] In one embodiment, the slider further includes a connecting neck connecting the slider body and the slider head, wherein the minimum lateral distance of the connecting neck perpendicular to the slider movement direction is d, and 0.5≤d / t≤0.73.

[0007] In one embodiment, the connecting neck and the slide body are smoothly connected by an arc surface.

[0008] In one embodiment, the radius of the arc surface is r, where 0.1 mm ≤ r ≤ 0.5 mm.

[0009] In one embodiment, 0.55 ≤ D1 / t ≤ 0.98; and / or, the surface roughness of the slider head is Ra, Ra ≤ 1.6 μm.

[0010] In one implementation, 2e / Lv ≤ 0.51.

[0011] In one implementation, 0.11 ≤ t / Lv ≤ 0.21.

[0012] In one embodiment, the slide is made of alloy steel.

[0013] In one embodiment, the alloy steel is configured as at least one of high-speed steel or stainless steel; And / or, the surface hardness value of the alloy steel is higher than HV700.

[0014] In one embodiment, the cylinder body is provided with at least one.

[0015] The present invention also proposes a refrigeration device, which includes the compressor described above.

[0016] In the technical solution of the present invention, the vane head is located in the receiving groove, so that the vane head is hinged to the roller as a whole. The vane body is located in the vane groove. Since the roller is eccentrically rotatable in the working chamber through the crankshaft, when the compressor is working, the vane body will extend out of the vane groove under the drive of the roller, and the movement stroke of the vane body relative to the vane groove is 2e, forming a variable compression chamber to realize the intake, compression and discharge of gas.

[0017] By establishing and defining the mathematical relationship between the key dimensions of the vane (total length Lv, thickness t, vane head outer diameter D1) and the core parameters of the compressor (eccentricity e, cylinder height H, cylinder body inner diameter D4), the following condition is met: 0.09 ≤ With a value of ≤2.2, it can systematically cope with the combined load generated by friction and gas force, optimize the strength and wear resistance of the vane, and thus improve the reliability and durability of the vane under complex working conditions, thereby improving the overall reliability of the compressor. Attached Figure Description

[0018] To more clearly illustrate the technical solutions in the embodiments of the present invention or the prior art, the drawings used in the description of the embodiments or the prior art will be briefly introduced below. Obviously, the drawings described below are only some embodiments of the present invention. For those skilled in the art, other drawings can be obtained based on the structures shown in these drawings without creative effort.

[0019] Figure 1 This is a schematic diagram of the structure of an embodiment of the compressor provided by the present invention; Figure 2 This is a schematic diagram of the cylinder body. Figure 3 for Figure 2 Cross-sectional view of the cylinder body; Figure 4 This is a schematic diagram of the slider structure; Figure 5 This is a schematic diagram of the crankshaft structure; Figure 6 This is a cross-sectional view of the compressor; Figure 7 This is a schematic diagram of the structure of the twin-cylinder pump body assembly in a twin-cylinder compressor. Figure 8 This is a trend graph showing the oil film thickness at the slider head.

[0020] Explanation of icon numbers: 1. Compressor; 10. Cylinder body; 11. Working chamber; 12. Vane groove; 20. Roller; 21. Receiving groove; 30. Slider; 31. Slider body; 32. Slider head; 33. Connecting neck; 34. Curved surface; 40. Crankshaft.

[0021] The realization of the objective, functional features and advantages of the present invention will be further explained in conjunction with the embodiments and with reference to the accompanying drawings. Detailed Implementation

[0022] The technical solutions of the embodiments of the present invention will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only a part of the embodiments of the present invention, and not all of the embodiments. Based on the embodiments of the present invention, all other embodiments obtained by those of ordinary skill in the art without creative effort are within the scope of protection of the present invention.

[0023] It should be noted that if the embodiments of the present invention involve directional indicators (such as up, down, left, right, front, back, etc.), the directional indicators are only used to explain the relative positional relationship and movement of the components in a specific posture. If the specific posture changes, the directional indicators will also change accordingly.

[0024] Furthermore, if the embodiments of this invention involve descriptions such as "first" or "second," these descriptions are for descriptive purposes only and should not be construed as indicating or implying their relative importance or implicitly specifying the number of technical features indicated. Thus, a feature defined with "first" or "second" may explicitly or implicitly include at least one of those features. Additionally, the use of "and / or" or "and / or" throughout the text includes three parallel solutions. For example, "A and / or B" includes solution A, solution B, or a solution where both A and B are satisfied simultaneously. Furthermore, the technical solutions of the various embodiments can be combined with each other, but this must be based on the ability of those skilled in the art to implement them. When the combination of technical solutions is contradictory or impossible to implement, it should be considered that such a combination of technical solutions does not exist and is not within the scope of protection claimed by this invention.

[0025] Roller compressors rely on eccentric rollers and the cylinder to form a changing working chamber to compress gas. Sliding vanes reciprocate within vane grooves to separate high and low pressure chambers and maintain a seal. However, existing vane designs often have inadequate key geometric parameters (such as head diameter, overall length, and thickness) that fail to match the eccentricity and displacement. This can easily lead to vane failure due to stress concentration, excessive wear, or fatigue, consequently affecting the compressor's sealing performance, efficiency, and service life.

[0026] During operation, the frictional force on the side of the vane is positively correlated with the crankshaft eccentricity, while the gas force on its head and sides is positively correlated with the compressor displacement. The larger the product of eccentricity and displacement, the higher the overall load on the vane, and the more stringent the requirements for its reliability.

[0027] To address this technical problem, the present invention proposes a compressor.

[0028] Please see Figures 1 to 5In one embodiment of the present invention, the compressor 1 includes a cylinder body 10, rollers 20, a crankshaft, and vanes 30. The cylinder body 10 has a working chamber 11 and a vane groove 12 communicating with the working chamber 11. The rollers 20 are eccentrically rotatable within the working chamber 11 via the crankshaft, and the outer peripheral wall of the rollers 20 has a recessed receiving groove 21 that extends along its axial direction. The vane 30 includes a vane body 31 and a vane connected to each other. The head 32, the slide body 31 is reciprocating within the slide groove 12, and the slide head 32 is adapted to be installed and hinged within the receiving groove 21; the total radial length of the slide 30 is defined as Lv, the thickness of the slide 30 extending circumferentially within the cylinder body 10 is t, the outer diameter of the slide head 32 is D1, the eccentricity of the crankshaft is e, and the cylinder height and inner diameter of the cylinder body 10 are H and D4 respectively, satisfying: 0.09≤[10 -3 *e*H*π*((D4)^2-(D4-2e)^2)] / (4*Lv* )≤2.2; This can constrain the reliability of the vane 30 to cope with the combined load of friction and gas force generated therefrom, optimize the strength, wear resistance and motion stability of the vane 30, thereby improving the overall reliability of the compressor 1.

[0029] In the technical solution of the present invention, the vane head 32 of the vane 30 is located in the receiving groove 21, so that the vane head 32 and the roller 20 are hinged into a whole. The vane body 31 is disposed in the vane groove 12. Since the roller 20 is eccentrically rotatably disposed in the working chamber 11 through the crankshaft, when the compressor 1 is working, the vane body 31 will extend out of the vane groove 12 under the drive of the roller 20, and the movement stroke of the vane body 31 relative to the vane groove 12 is 2e, forming a variable compression chamber to realize the intake, compression and discharge of gas.

[0030] By establishing and defining the mathematical relationship between the key dimensions of the vane 30 (total length Lv, thickness t, outer diameter D1 of the vane head 32) and the core parameters of the compressor 1 (eccentricity e, cylinder height H, inner diameter D4 of the cylinder body 10), the following condition is met: 0.09 ≤ [10 -3 *e*H*π*((D4)^2-(D4-2e)^2)] / (4*Lv* With a strength of ≤2.2, it can systematically cope with the combined load generated by friction and gas force, optimize the strength and wear resistance of the vane 30, and thus improve the reliability and durability of the vane 30 under complex working conditions, thereby improving the overall reliability of the compressor 1.

[0031] It should be noted that the key dimensions of the vane 30 directly affect its strength, rigidity, and frictional characteristics; while the core parameters of the compressor 1 determine the magnitude of the mechanical friction and gas forces borne by the vane 30. The magnitude of the eccentricity is positively correlated with the frictional force on the side of the vane 30, that is, the larger the eccentricity, the more intense the lateral movement of the vane 30, leading to an increase in side friction. The displacement is positively correlated with the gas forces on the vane head 32 and the side of the vane 30, that is, the higher the displacement and pressure difference, the greater the gas forces on the vane head 32 and the side. These two types of loads together constitute the comprehensive load of the vane 30.

[0032] Specifically, the slider 30 includes a slider head 32 and a slider body 31 connected together. Since the slider groove 12 extends axially along the cylinder body 10 and radially along the cylinder body 10, the slider body 31 is adapted to the slider groove 12, resulting in a rectangular or approximately rectangular cross-section. The slider head 32 is hinged to the receiving groove 21, which is specifically an arc groove. In a cross-section perpendicular to the roller 20 axis, the profile of the receiving groove 21 is an arc shape (wrap angle > 180°). The slider head 32 is adapted to contact the receiving groove 21, giving it a minimum circumscribed circle diameter D1. The length Lv is the total radial length of the slider 30 in the cylinder body 10, and the thickness t is the thickness of the slider 30 extending circumferentially in the cylinder body 10, which is also the thickness of the slider body 30. The unit is mm. These parameters can be measured using tools such as calipers and optical profilometers.

[0033] And the relation S:

[10] -3 *e*H*π*((D4)^2-(D4-2e)^2)] / (4*Lv* ) actually composed of [10 -3 *e*H*π*((D4)^2-(D4-2e)^2)] / 4 divided by (Lv* ) is obtained, where, through [10 -3 *H*π*((D4)^2-(D4-2e)^2)] / 4 determines the theoretical displacement level of compressor 1, e*[10 -3 *H*π*((D4)^2-(D4-2e)^2)] / 4 is the product of displacement and eccentricity, which represents the load on the high-pressure working chamber within compressor 1. A larger product indicates a larger load on the high-pressure working chamber, requiring higher reliability of the sliding vane 30. Through (Lv* The combination of key geometric dimensions of the slider 30 is controlled to ensure that the slider 30 has sufficient structural strength and rigidity to reliably withstand the complex load generated by the combined action of gas pressure and friction, thus determining the reliability of the slider 30.

[0034] Through

[10] -3*e*H*π*((D4)^2-(D4-2e)^2)] / (4*Lv* The displacement requirement is systematically matched with the load-bearing capacity of the vane 30 to ensure long-term reliable operation of the vane 30 under harsh conditions such as high displacement and high eccentricity. When S is too large, i.e., greater than 2.2, it means pursuing high displacement with relatively small vane 30 structural dimensions. This will lead to stress concentration at the root of the vane 30, excessively high contact surface specific pressure, and aggravated bending deformation, thus significantly increasing the risk of jamming, breakage, or abnormal wear. Conversely, when S is too small, i.e., less than 0.09, a larger t, D1, or Lv is used to support a smaller displacement. Although this provides sufficient safety margin and extremely reliable operation, it will easily increase the mass of the vane 30, leading to increased inertial force and aggravated vibration and noise; it will also increase the contact area between the vane 30 and the vane groove 12, increasing friction loss and reducing volumetric efficiency and energy efficiency; at the same time, it is not conducive to the miniaturization and weight reduction of the compressor 1, and increases costs.

[0035] Therefore, by limiting

[10] -3 *e*H*π*((D4)^2-(D4-2e)^2)] / (4*Lv* Within the range of 0.09 to 2.2, the geometry of the vane 30 (thickness t, vane head 32 diameter D1, and length Lv) is well-coordinated with the displacement level of the compressor 1. At this point, the vane 30 can effectively withstand the combined load of gas force and friction without increasing unnecessary volume or mass due to structural redundancy. This facilitates achieving good sealing performance, low wear rate, and high mechanical efficiency, while also ensuring a compact structure and manufacturing economy.

[0036] Among them, [10 -3 *e*H*π*((D4)^2-(D4-2e)^2)] / (4*Lv* The specific values ​​of ) include, but are not limited to, 0.09, 0.095, 0.1, 0.15, 0.2, 0.25, 0.3, 0.35, 0.37, 0.4, 0.45, 0.5, 0.55, 0.6, 0.65, 0.7, 0.75, 0.8, 0.95, 1, 1.05, 1.35, and 2.2. However, in other embodiments, while ensuring the performance of compressor 1,

[10] -3 *e*H*π*((D4)^2-(D4-2e)^2)] / (4*Lv* It can be greater than 2.2 or less than 0.09.

[0037] Oil film thickness is a key indicator affecting the lubrication condition and operational reliability of the compressor's vane assembly; a thicker oil film results in better vane reliability. Please refer to [link / reference needed]. Figure 8In embodiments of the present invention, it is further defined that: 0.37 ≤

[10] -3 *e*H*π*((D4)^2-(D4-2e)^2)] / (4*Lv* With a thickness of ≤1.05, the oil film thickness can be further maintained within the effective lubrication range of 0.1μm-2μm, effectively avoiding direct contact between components, reducing wear, frictional heat, and fretting fatigue risks, and improving the overall operational reliability and service life of the machine. At this point, while ensuring sufficient structural strength and rigidity, the sliding vane 30 also ensures good coordination between its geometric dimensions (thickness t, vane head diameter D1, and length Lv), the compressor displacement level, and lubrication performance.

[0038] Please see Figure 4 In an embodiment of the present invention, the slider 30 further includes a connecting neck 33 connecting the slider body 31 and the slider head 32. The minimum lateral distance of the connecting neck 33 perpendicular to the direction of movement of the slider 30 is d, where 0.5 ≤ d / t ≤ 0.73. The thickness t at this point determines the circumferential cross-sectional dimensions of the slider 30, thereby determining the contact area between the slider 30 and the sidewall of the slider groove 12, the magnitude of friction, and the bending resistance in the radial plane. The lateral distance d controls the effective load-bearing cross-sectional width of the connecting neck 33 in the circumferential plane. d / t reflects the structural continuity and strength concentration of the slider 30 within the same circumferential cross-section.

[0039] When d / t is too large, i.e., greater than 0.73, the lateral distance of the connecting neck 33 is prone to be too large, which restricts the hinge flexibility of the slider head 32 and also causes material redundancy, which is not conducive to weight reduction. When d / t is too small, i.e., less than 0.5, the connecting neck 33 forms a significant necking, becoming a weak link in the slider 30 as a whole. Since the slider 30 is subjected to alternating high-pressure gas force and reciprocating inertial force during operation, this area is prone to high stress concentration, inducing microcracks and rapidly propagating them, ultimately leading to fracture failure.

[0040] Therefore, by limiting d / t to the range of 0.5 and 0.73, the connecting neck 33 can retain a sufficient effective cross-section, improving its structural strength. This effectively transmits gas forces and inertial loads while avoiding excessive stress concentration. In this configuration, the slider 30 exhibits good fatigue resistance during high-frequency reciprocating motion, while allowing for necessary functional optimization of the slider head 32 (such as rounded corner transitions or localized weight reduction), balancing strength, lightweight design, and manufacturing feasibility.

[0041] The specific values ​​of d / t include, but are not limited to, 0.5, 0.55, 0.6, 0.65, 0.7, and 0.73. However, in other embodiments, while ensuring the performance of compressor 1, d / t can be greater than 0.73 or less than 0.5.

[0042] Optionally, in an embodiment of the present invention, the connecting neck 33 and the slider body 31 are smoothly connected by an arc surface 34. The arc surface 34 is tangent at the connection point between the connecting neck 33 and the slider body 31, forming a continuous curved surface without sharp corners. That is, a rounded corner treatment is used at the transition point between the connecting neck 33 and the slider body 31 to reduce stress concentration, thereby improving the overall reliability and durability of the slider 30. The rounded corner transition design not only optimizes the force transmission path but also effectively avoids local stress peaks caused by right-angle connections. It also makes high-precision machining easier to achieve, ensuring the stability and reliability of the slider 30 in high-frequency reciprocating motion.

[0043] Specifically, in an embodiment of the present invention, the radius of the arc surface 34 is r, where 0.1 mm ≤ r ≤ 0.5 mm. In practical applications, the specific value of the radius of the arc surface 34 can be flexibly adjusted according to the material of the slider 30, the working environment, and the requirements of the manufacturing process.

[0044] When the radius of the arc surface 34 is too small, a sharp geometric abrupt change will occur in the transition area, resulting in severe stress concentration under high-frequency alternating loads. This can easily induce microcracks and accelerate fatigue fracture. At the same time, small fillets are more sensitive to surface defects (such as tool marks and burrs), further reducing reliability. Conversely, when the fillet radius is too large, although it can improve stress distribution, it will excessively erode the effective load-bearing section of the connecting neck 33, weakening the structural strength. This may exacerbate the insufficient strength problem, especially when the d / t ratio is already small. In addition, excessively large fillets may also encroach on the functional area of ​​the slider head 32, restrict the overall layout, and increase the difficulty and cost of precision machining.

[0045] Therefore, controlling the radius of the arc surface 34 within the range of 0.1mm to 0.5mm significantly alleviates stress concentration and improves fatigue life while retaining sufficient load-bearing area, and is easily achievable through conventional grinding or wire cutting processes. Combined with a reasonable d / t ratio, this fillet design effectively optimizes the stress distribution of the slider 30 in the main load-bearing plane (circumferential-radial plane), thus balancing the strength, durability, and manufacturability of the slider 30. Specific values ​​for the radius of the arc surface 34 include, but are not limited to, 0.1mm, 0.2mm, 0.3mm, 0.4mm, and 0.5mm.

[0046] Optionally, in an embodiment of the present invention, 0.55≤D1 / t≤0.98; D1 / t is the ratio of the diameter D1 of the slider head 32 to the thickness t of the slider 30. When D1 / t is too large, i.e. greater than 0.98, the slider head 32 has a larger volume and higher mass, which will increase the inertial force of the slider 30 reciprocating motion, thereby aggravating the impact and wear on the side wall of the slider groove 12, especially under high speed conditions. In addition, an excessively large slider head 32 may also restrict the degree of freedom of movement of the slider 30 in extreme eccentric positions, and may even cause motion interference with the roller 20 or the cylinder.

[0047] Conversely, when D1 / t is too small, i.e., less than 0.55, the size of the slider head 32 is too small, making it difficult to provide sufficient structural strength and load-bearing area. This not only easily leads to stress concentration under alternating loads, causing crack initiation or even fracture, but also reduces the area of ​​contact with the roller 20, resulting in a significant increase in contact stress and exacerbating the risks of wear and crushing. Simultaneously, it is detrimental to the lubrication and heat dissipation of the friction pair, affecting the motion reliability and service life of the hinged pair between the slider 30 and the roller 20.

[0048] Therefore, by limiting D1 / t to the range of 0.55 and 0.98, not only is the load-bearing size of the vane head 32 optimized, but it also better ensures surface quality during processing. A reasonable vane head 32 size provides sufficient hinge contact area, effectively reducing contact stress per unit area, thereby reducing wear between the roller 20 and the vane head 32 and extending the overall service life of the compressor 1. Simultaneously, a suitable geometric proportion facilitates precision grinding or lapping with standard tools, achieving higher surface finish and dimensional accuracy, further improving the mating performance of the friction pair. Furthermore, good surface quality contributes to the formation and maintenance of the lubricating oil film, enhancing wear resistance and fatigue resistance during high-frequency reciprocating motion.

[0049] The specific values ​​of D1 / t include, but are not limited to, 0.55, 0.6, 0.65, 0.7, 0.73, 0.8, 0.85, 0.9, and 0.98. However, in other embodiments, while ensuring the performance of compressor 1, D1 / t can be greater than 0.98 or less than 0.55.

[0050] Specifically, in an embodiment of the present invention, the surface roughness of the slide head 32 is Ra, where Ra ≤ 1.6 μm. This level of precision can be achieved through final processing methods such as ultra-fine grinding, lapping, or polishing. The surface roughness of the slide head 32 not exceeding 1.6 μm not only reduces the height of the microscopic contact peaks between the slide head 32 and the roller 20, thereby reducing the coefficient of friction and starting / running resistance; it also facilitates the formation of a stable oil film under limited lubrication conditions or boundary lubrication conditions, reducing direct metal-to-metal contact and effectively suppressing adhesive wear, abrasive wear, and micro-pitting; simultaneously, it alleviates local stress concentration, delays the initiation of fatigue cracks, and improves the fatigue life of components under high-frequency alternating loads.

[0051] Optionally, in an embodiment of the present invention, 2e / Lv ≤ 0.51. The total length Lv of the slider 30 needs to be designed according to the eccentricity e. By limiting the ratio of the crankshaft eccentricity e to the total length Lv of the slider 30, the relationship between the effective working stroke of the slider 30 and its overall support length during reciprocating motion is reflected, which is used to characterize the degree of motion eccentricity of the slider 30 system, referred to as the eccentricity ratio.

[0052] When the eccentricity is too large, the overhang length of the slide vane 30 increases, leading to a decrease in bending stiffness. Under the action of gas force and inertial force, it is prone to large flexural deformation, which in turn aggravates the contact stress and frictional wear with the sidewall of the slide vane groove 12. When the eccentricity is too small, it will lead to structural redundancy and reduced displacement efficiency, which is not conducive to compact design. Limiting the eccentricity to no more than 0.51 can meet the displacement requirements while maintaining good bending stiffness and smooth movement, effectively suppressing the second-order oscillation of the slide vane 30, ensuring the continuity of the sealing line, and reducing leakage and frictional loss. Combined with the synergistic design of optimized slide vane head 32 dimensions, surface roughness control, application of high-hardness materials, and adequate lubrication, the durability of the slide vane 30-slide vane groove 12 pair and the hinge pair can be further improved, thereby achieving high reliability, low noise, and long service life under high-frequency alternating loads.

[0053] Optionally, in embodiments of the present invention, 0.11 ≤ t / Lv ≤ 0.21. When t / Lv is too large, i.e. greater than 0.21, the mass of the slide vane 30 increases, and the inertial force rises, which not only puts an extra burden on the drive system, but also occupies more internal space in the cylinder, limiting the optimization of eccentricity or roller 20 size. At the same time, the material usage and processing cost also increase. When t / Lv is too small, i.e. less than 0.11, the rigidity of the slide vane 30 is insufficient, and it is prone to bending deformation under high pressure differential and high-speed reciprocating motion, leading to sealing failure, increased sidewall wear, or even jamming.

[0054] Therefore, by limiting t / Lv to the range of 0.11 and 0.21, it is possible to avoid over-design while ensuring sufficient bending stiffness and fatigue strength, effectively control material consumption, heat treatment difficulty and precision machining cost, thereby achieving a balance between structural strength, motion reliability and manufacturing cost.

[0055] The specific values ​​of t / Lv include, but are not limited to, 0.11, 0.12, 0.13, 0.14, 0.15, 0.16, 0.17, 0.18, 0.19, 0.2, and 0.21. However, in other embodiments, while ensuring the performance of compressor 1, t / Lv can be greater than 0.21 or less than 0.11.

[0056] Furthermore, in an embodiment of the present invention, the slide 30 is made of alloy steel. It possesses excellent wear resistance and fatigue resistance, enabling it to withstand the mechanical stresses caused by high pressure differentials and high-speed reciprocating motion. In addition, after appropriate heat treatment, the hardness and toughness of the alloy steel material can be further improved, thereby extending the service life of the slide 30 and ensuring stable performance even under extreme operating conditions.

[0057] Optionally, in embodiments of the present invention, the alloy steel is configured as at least one of high-speed steel or stainless steel; and / or, the surface hardness value of the alloy steel is higher than HV700. Thus, the slide 30 is made of high-speed steel and / or stainless steel, and its surface hardness is further improved through a suitable heat treatment process to ensure that it exceeds the HV700 standard. This material selection and processing can significantly improve the wear resistance of the slide 30, thereby extending its service life and enhancing its operational stability.

[0058] The optimized series of dimensional ratios described above ensure that the vane 30 has sufficient neck strength, excellent head wear resistance, and stable motion characteristics under complex load conditions, thereby significantly improving the reliability and durability of the articulated roller compressor 1 under high load conditions.

[0059] Please see Figure 6 and Figure 7 In embodiments of the present invention, at least one cylinder body 10 is provided. The number and arrangement of cylinder bodies 10 can be adjusted according to actual needs to meet the usage requirements under different working conditions. By rationally designing the cylinder layout, the overall structural compactness of the compressor 1 can be effectively optimized, while improving its operating efficiency.

[0060] Specifically, compressor 1 can be configured as a single-cylinder compressor 1. In this case, the single-cylinder compressor 1 includes a cylinder body 10, a crankshaft with an eccentric portion, rollers 20 mounted on the eccentric portion, and vanes 30 corresponding to the rollers 20, which are reciprocatingly slidably disposed in vane grooves 12 of the cylinder body 10. The vane head 32 is always in hinged contact with the roller 20, thereby forming two periodically changing working chambers in the cylinder cavity to realize the intake, compression, and exhaust processes. This structure can also be expanded to a dual-cylinder or multi-cylinder compressor 1. In a dual-cylinder compressor 1, the crankshaft typically has two eccentric portions with a 180° phase difference, located at different axial positions. Each eccentric portion is fitted with a roller 20, and each corresponds to an independent cylinder body 10 and a matching vane 30. The two compression units share the same crankshaft, and the phase misalignment effectively reduces torque fluctuations and improves operational stability. For multi-cylinder (e.g., three-cylinder or higher) compressors 1, the number of eccentric sections and cylinder units is further increased, typically used in large-displacement or high-power applications such as commercial refrigeration, heat pumps, or air conditioning systems. Regardless of whether it is a single-cylinder, dual-cylinder, or multi-cylinder configuration, the core working principle is the same, relying on the coordinated operation of eccentric rotary motion and reciprocating motion of the vane 30 to achieve continuous and stable gas compression. Furthermore, the specific dimensional relationships of each cylinder body 10 and its corresponding rollers 20, vanes 30, etc., within the compressor 1 can be specifically defined by referring to the above embodiments.

[0061] The present invention also proposes a refrigeration device, which includes a compressor 1. The specific structure of the compressor 1 is as described in the above embodiments. Since the refrigeration device adopts all the technical solutions of all the above embodiments, it has at least all the beneficial effects brought about by the technical solutions of the above embodiments, which will not be described in detail here.

[0062] The above description is merely an exemplary embodiment of the present invention and does not limit the scope of protection of the present invention. Any equivalent structural transformations made based on the technical concept of the present invention and the contents of the specification and drawings of the present invention, or direct / indirect applications in other related technical fields, are included within the scope of protection of the present invention.

Claims

1. A compressor, characterized in that, include: The cylinder body has a working chamber inside, and the cylinder body also has a sliding vane groove that communicates with the working chamber; A roller and a crankshaft, wherein the roller is eccentrically rotatable within the working cavity via the crankshaft, and the outer peripheral wall of the roller is recessed with a receiving groove extending along its axial direction; A slider includes a slider body and a slider head connected to each other. The slider body can reciprocate within the slider groove, and the slider head is adapted to be installed and hinged within the receiving groove. Define the total radial length of the slide vane as Lv, the thickness of the slide vane extending circumferentially within the cylinder body as t, the outer diameter of the slide vane head as D1, the eccentricity of the crankshaft as e, and the cylinder height and inner diameter of the cylinder body as H and D4, respectively, satisfying: 0.09≤ ≤2.2。 2. The compressor as described in claim 1, characterized in that, 0.37≤ ≤1.05。 3. The compressor as described in claim 1, characterized in that, The slider also includes a connecting neck that connects the slider body and the slider head. The minimum lateral distance of the connecting neck perpendicular to the slider movement direction is d, where 0.5 ≤ d / t ≤ 0.

73.

4. The compressor as described in claim 3, characterized in that, The connecting neck and the slide body are smoothly connected by an arc surface.

5. The compressor as described in claim 4, characterized in that, The radius of the arc surface is r, where 0.1mm ≤ r ≤ 0.5mm.

6. The compressor as claimed in claim 1, characterized in that, 0.55≤D1 / t≤0.98; And / or, the surface roughness of the slider head is Ra, Ra≤1.6μm.

7. The compressor as claimed in claim 1, characterized in that, 2e / Lv≤0.

51.

8. The compressor as claimed in claim 1, characterized in that, 0.11≤t / Lv≤0.

21.

9. The compressor as claimed in claim 1, characterized in that, The slider is made of alloy steel.

10. The compressor as claimed in claim 9, characterized in that, The alloy steel is configured to be at least one of high-speed steel or stainless steel; And / or, the surface hardness value of the alloy steel is higher than HV700.

11. The compressor as claimed in claim 1, characterized in that, The cylinder body is provided with at least one.

12. A refrigeration device, characterized in that, Includes the compressor as described in any one of claims 1 to 11.

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