Spring bushing for small refrigeration compressor

Abstract

The utility model discloses a spring bushing for a small refrigeration compressor, which comprises a double-layer bushing, and an inner cavity of the double-layer bushing is matched with a convex nail of a lower shell of the compressor to form interference fit; the double-layer lining comprises an outer rod body and an inner rod body which are coaxially arranged, the outer rod body is sleeved with an outer spring, and the inner rod body is sleeved with an inner spring; the outer spring stiffness is smaller than the inner spring stiffness. An integrated double-layer bushing structure is formed by the outer rod body and the inner rod body which are coaxially nested, so that the problem of contradiction between resonance avoidance and horizontal displacement control caused by spring stiffness difference of a small compressor is solved, the collaborative optimization effect of vertical vibration energy absorption and horizontal swing suppression is achieved, the vibration acceleration of a compressor shell is reduced, and the vibration speed of the compressor shell is increased. Abnormal noise is reduced, and reliability of the whole machine is improved.

Description

Technical Field

[0001] This utility model relates to the field of compressor components, and in particular to a spring bushing for a small refrigeration compressor. Background Technology

[0002] With the accelerating trend of energy conservation and miniaturization in home appliances, inverter compressors, as a core component of refrigerators, are facing new technological challenges. During the miniaturization process, the natural frequency of traditional fixed-frequency compressors increases due to the reduction in the weight of the compressor core. Its resonant frequency band is prone to falling into the operating speed range of inverter compressors (1200-4500 r / min). While simply reducing the stiffness of the support springs can adjust the resonant frequency, it will cause abnormal horizontal vibration of the compressor core, which seriously restricts the reliability of the compressor.

[0003] In the existing technology, compressor vibration mainly originates from three aspects: first, the airflow pulsation caused by the change of working fluid physical parameters during the intake and exhaust process, which will lead to resonance when the pulsation frequency is coupled with the natural frequency of the pipeline; second, the mechanical vibration generated by the unbalanced force between the moving pairs of the crankshaft-connecting rod mechanism and the valve plate slapping; and third, the structural vibration caused by the alternating magnetic field of the electromagnetic components.

[0004] For example, in the Chinese patent document "An Aeration Tank for Treating Wastewater" (publication number CN112483353B), at least one set of support components is provided at the bottom of the compressor core. The support components consist of a spring, an upper spring base, and a lower spring base. The upper spring base is fixed to the bottom surface of the compressor core, and both ends of the spring are connected to the upper spring base and the lower spring base, respectively. This invention uses a support component at the bottom of the compressor core to reduce the amplitude of the compressor core's vibration. The spring structure of the support component is improved so that the spring can simultaneously meet the requirements of both high-frequency and low-frequency vibrations of the compressor on the support system, reducing vibrations during compressor start-up and shutdown. The spring's stiffness can change with the amplitude of vibrations during compressor operation, providing better absorption of smaller amplitude vibrations and better suppression of larger amplitude vibrations.

[0005] The above-described solution effectively reduces compressor operating vibration, increases compressor reliability, and positively contributes to extending compressor lifespan. However, while optimizing pipeline layout and improving dynamic balance design can alleviate vibration issues to some extent, conventional single-layer spring structures exhibit inherent limitations in addressing the contradiction between reduced spring stiffness and horizontal vibration suppression. When using low-stiffness springs to avoid resonance, the compressor core is prone to lateral displacement exceeding design limits during start-up, shutdown, and variable-speed operation. This phenomenon not only exacerbates wear on moving parts but also transmits abnormal noise through the casing, directly affecting the overall performance of the refrigerator. Therefore, it is urgent to develop a new elastic support structure that can effectively constrain the horizontal displacement of the compressor core while maintaining low-stiffness frequency regulation capabilities. Utility Model Content

[0006] To address the inherent limitations of conventional single-layer spring structures in addressing the contradiction between reduced spring stiffness and suppression of horizontal vibration in existing technologies, this invention utilizes a coaxially nested outer and inner rods to form an integrated double-layer bushing structure. This, combined with a layout where the outer rod fixes the outer spring and the inner rod constrains the inner spring, resolves the conflict between resonance avoidance and horizontal displacement control in small compressors caused by differences in spring stiffness. This achieves a synergistic optimization effect of vertical vibration energy absorption and horizontal sway suppression, reducing compressor casing vibration acceleration, decreasing abnormal noise, and improving overall machine reliability.

[0007] To achieve the above objectives, the present invention adopts the following technical solution:

[0008] A spring bushing for a small refrigeration compressor includes a double-layer bushing, the inner cavity of which is adapted to the protrusion of the compressor lower housing to form an interference fit; the double-layer bushing includes an outer rod and an inner rod arranged coaxially, an outer spring is sleeved on the outer rod, and an inner spring is sleeved on the inner rod.

[0009] Preferably, the outer rod has circumferentially distributed outer rod protrusions on its outer circumferential surface, and the inner rod has circumferentially distributed inner rod protrusions on its outer circumferential surface. The top heights of the outer rod protrusions and the inner rod protrusions respectively extend beyond the corresponding outer circumferential surface of the rod.

[0010] Preferably, the cross-sectional shapes of the outer rod protrusion and the inner rod protrusion are arc-shaped or trapezoidal, and their shapes may be the same or different.

[0011] Preferably, the outer rod and the outer spring, and the inner rod and the inner spring, are respectively fitted with a clearance fit or an interference fit.

[0012] Preferably, the bottom end face of the double-layer bushing forms a planar contact with the mounting base surface of the protruding nail, and an annular reinforcing rib is provided at the connection between the outer rod and the inner rod.

[0013] Preferably, the outer rod protrusions are distributed in a spiral array along the axial direction of the outer rod body, and the inner rod protrusions are distributed in a straight array with equal spacing along the axial direction of the inner rod body.

[0014] Preferably, the outer circumferential diameter of the outer rod is larger than the outer circumferential diameter of the inner rod, and the difference between the two diameters is 1.1-3 times the difference in wire diameter between the outer spring and the inner spring.

[0015] Preferably, the wall thickness of the double-layer bushing gradually decreases from the bottom to the top, and the bottom wall thickness is 1.1-3 times the top wall thickness.

[0016] Preferably, the outer rod and inner rod of the double-layer bushing are integrally formed.

[0017] Therefore, this utility model has the following beneficial effects:

[0018] The system's natural frequency is reduced by using an outer low-stiffness spring, causing the resonance threshold to be removed from the compressor's operating speed range; the inner high-stiffness spring forms a lateral displacement constraint ring, controlling the horizontal sway amplitude of the mechanism within the design threshold, thus avoiding structural fatigue caused by abnormal vibration.

[0019] Low-stiffness springs preferentially absorb vertical vibration energy, while high-stiffness springs primarily suppress horizontal displacement. The two are rigidly connected by bushings to isolate the vibration transmission path, reducing the vibration acceleration of the compressor housing by more than 40% and effectively blocking the noise transmission channel.

[0020] The inner and outer rods of the double-layer bushing adopt an interference fit and irregular protrusion combination design, which not only ensures the spring installation and positioning accuracy, but also consumes high-frequency vibration energy through friction damping, thereby reducing the deviation of the spring axial preload by 60% and avoiding the dynamic instability caused by traditional clearance fit.

[0021] The one-piece injection-molded double-layer bushing is 65% lighter than that of a metal bushing. The coaxial design of the inner cavity makes the stress distribution more uniform, and the improved oil resistance of the elastomer material significantly extends the overall service life, meeting the long-term reliable operation requirements of small compressors in compact spaces. Attached Figure Description

[0022] Figure 1 This is a schematic diagram of the structure of this utility model.

[0023] Figure 2 yes Figure 1 A schematic diagram of the structure of the double-layer bushing.

[0024] Figure 3 yes Figure 1 Top view of the double-layer bushing.

[0025] Figure 4 yes Figure 1 A magnified view of a portion of point A in the middle.

[0026] Figure 5 yes Figure 1 Cross-sectional view at point BB.

[0027] In the diagram: 100, double-layer bushing; 1, outer bushing rod; 11, annular reinforcing rib; 2, inner bushing rod; 3, outer rod protrusion; 4, inner rod protrusion; 5, compressor lower housing; 6, protruding nail; 7, outer spring; 8, inner spring. Detailed Implementation

[0028] The present invention will be further described below with reference to the accompanying drawings and specific embodiments.

[0029] Example 1

[0030] like Figure 1 As shown, a spring bushing for a small refrigeration compressor has the following specific structure. Figures 1 to 3 As shown, the device includes a double-layer bushing 100, an outer spring 7, and an inner spring 8. The double-layer bushing 100 is manufactured using an integral injection molding process, and the material is nylon. Its inner cavity forms an interference fit with the protruding pins 6 of the compressor lower housing 5. The double-layer bushing 100 is composed of an outer rod 1 and an inner rod 2 arranged coaxially. The overlap of the axes of the outer rod 1 and the inner rod 2 is controlled by a precision mold to ensure their coaxial relationship. The outer circumferential diameter of the outer rod 1 is larger than that of the inner rod 2. The difference in diameter between the two is set based on the difference in wire diameter between the outer spring 7 and the inner spring 8, forming a nested structure that adapts to different spring stiffnesses.

[0031] like Figure 2 , 3 As shown, in this embodiment, the outer rod 1 and the inner rod 2 are connected by annular reinforcing ribs 11. The reinforcing ribs are located at the connection point, and their radial width matches the bushing wall thickness, forming a continuous stress distribution structure. The outer circumferential surface of the outer rod 1 has uniformly distributed outer rod protrusions 3, with arc-shaped cross-sections and a top height exceeding the outer circumferential surface of the outer rod 1 by 0.8-1.2 mm. The outer circumferential surface of the inner rod 2 has inner rod protrusions 4, whose cross-sectional shape is consistent with the outer rod protrusions 3, and whose top height exceeds the outer circumferential surface of the inner rod 2 by 0.6-1.0 mm. The protrusion array is spirally distributed along the axial direction, with the spacing between adjacent protrusions being 1.1-3 times the spring wire diameter.

[0032] like Figure 4 , 5 As shown, the outer spring 7 is sleeved on the outer circumference of the outer rod 1, with its inner diameter forming a clearance fit with the outer circumferential diameter of the outer rod 1, the clearance being 0.1-0.3 mm. The inner spring 8 is sleeved on the outer circumference of the inner rod 2, and the fit between its inner diameter and the outer circumferential diameter of the inner rod 2 can be either clearance or interference fit. The outer rod protrusion 3 and the outer spring 7 are engaged with a helical gap; when the spring is compressed and moves downward, the protrusion engages between adjacent helical coils, restricting the horizontal displacement of the spring. The inner rod protrusion 4 forms the same constraint mechanism for the inner spring 8.

[0033] The bottom end face of the double-layer bushing 100 is a flat structure, which fits perfectly with the mounting base of the protruding nail 6. The interference fit between the inner cavity of the bushing and the protruding nail 6 is controlled within 0.05-0.1mm to ensure no axial movement after assembly. An annular groove is provided at the connection between the outer rod 1 and the inner rod 2. The groove depth is 1 / 3 of the wall thickness to accommodate the weld line during injection molding and to avoid stress concentration.

[0034] When the compressor starts, the internal mechanism operates at a speed of 1200-4500 r / min driven by a variable frequency motor, generating multi-directional vibration excitation: the outer spring 7 preferentially absorbs vertical vibration energy. Its low stiffness characteristic reduces the system's natural frequency below the operating speed range, avoiding resonance. The helical structure of the spring generates nonlinear damping during compression, attenuating the vibration acceleration in the 20-200Hz frequency band by more than 40%. The inner spring 8, through mechanical interlocking between the inner rod protrusion 4 and the spring helix, restricts the horizontal displacement of the internal mechanism. When the internal mechanism is subjected to centrifugal force and undergoes lateral displacement, the radial stiffness of the inner spring 8 (3-5 times higher than that of the outer spring) forms a rigid constraint ring, controlling the displacement within ±0.5mm. The stiffness difference between the inner and outer springs creates graded vibration reduction: the outer spring bears the main vertical vibration, while the inner spring is mainly used for horizontal limiting.

[0035] Under variable speed conditions, the coaxial design of the outer rod 1 and the inner rod 2 ensures the consistency of the vibration transmission path. When the movement experiences compound vibration, the staggered layout of the outer rod protrusion 3 and the inner rod protrusion 4 (outer protrusion helical angle 15°, inner protrusion linear array) disrupts the in-phase superposition of vibration waves, reducing the vibration energy transmitted to the housing by more than 30%.

[0036] In this embodiment, the wall thickness of the outer rod 1 gradually decreases from the bottom to the top, which, combined with the creep characteristics of nylon material, makes the stress distribution uniform. The annular reinforcing rib undergoes reverse deformation when the temperature changes, which offsets the loss of preload caused by the thermal expansion of the material, ensuring that the spring pressing force fluctuates by less than 10% under operating conditions from -20℃ to 80℃.

[0037] During the actual assembly of the compressor, the operator needs to vertically press the double-layer bushing 100 onto the protruding pin 6 of the lower housing 5. A special guide fixture is used during pressing to ensure that the axial deviation between the inner cavity of the bushing and the protruding pin 6 is less than 0.1mm. The outer rod protrusion 3 on the outer circumference of the outer rod body 1 adopts an arc cross-section design with an arc radius R1.0mm, forming a three-point contact with the helical gap of the outer spring 7. When the spring is compressed and moves downward, the top of the protrusion embeds between adjacent helical coils, generating a radial constraint force that limits the horizontal displacement of the spring to within ±0.8mm. The inner rod protrusion 4 of the inner rod body 2 adopts a trapezoidal cross-section with a top width of 0.8mm and a bottom width of 1.5mm, forming an interference fit with the wire diameter of the inner spring 8. The interference is controlled within 0.05-0.1mm to ensure that the spring does not slip circumferentially under high-speed vibration.

[0038] During the initial startup of the compressor, the outer spring 7 absorbs over 80% of the vertical impact energy through its nonlinear stiffness characteristics, reducing the system's natural frequency from 450Hz to 280Hz, thus avoiding the 20-75Hz excitation frequency band corresponding to the operating speed of 1200-4500 r / min. When the speed jumps to 3000 r / min, the inner spring 8 generates a radial restoring force of 12-15N through the wedge-shaped action of the trapezoidal inclined surface of the inner rod protrusion 4 and the spring coil, effectively suppressing the horizontal sway of the compressor core caused by centrifugal force. Experimental data shows that this structure reduces the vibration intensity of the compressor casing at 1 / 3 octave band from 4.5 mm / s to 1.2 mm / s, and reduces the peak noise level by 8 dB(A).

[0039] The thickened area at the bottom of the bushing, with its interference fit to the protruding stud 6, forms a rigid support surface, capable of withstanding axial impact loads of up to 1500N during compressor start-up and shutdown. The thin-walled area at the top absorbs high-frequency vibrations through elastic deformation, preventing stress concentration and creep of the nylon material. After 200 hours of durability testing, the maximum deformation of the bushing is less than 0.05mm, and the spring preload decay rate is less than 5%, meeting the 10-year design life requirement.

[0040] Example 2

[0041] In this embodiment, another type of spring bushing for a small refrigeration compressor is provided. Its overall structure shares the same core features as the double-layer bushing 100 described in Embodiment 1, including a coaxially nested outer rod 1 and inner rod 2, and the installation method is the same. Specifically, the structures of the outer rod protrusion 3 and the inner rod protrusion 4 differ in this embodiment. The outer rod protrusion 3 is arranged axially on the outer circumferential surface of the outer rod 1, with the protrusion array arranged in a single-line spiral distribution. The spiral helix angle is 18°±2°, and the axial spacing between adjacent protrusions is 1.8-2.2 times the wire diameter of the outer spring 7. Each protrusion has a trapezoidal cross-section, with a top width of 1.2mm, a bottom width of 2.0mm, a height of 1.5mm, and a trapezoidal slope angle of 45°. The spiral direction is opposite to the rotation direction of the compressor core; when the core rotates clockwise, the protrusion spiral is left-handed. The inner rod protrusions 4 on the outer circumferential surface of the inner rod 2 are arranged in an axially equidistant straight line, with the spacing between adjacent protrusions precisely equal to the pitch of the inner spring 8. The raised section is semi-circular with a radius of 0.8 mm, and its top centerline is aligned with the spring helical gap. The linear array extends to 2 / 3 of the total length of the inner rod 2, while the bottom 1 / 3 is a smooth surface used to accommodate the spring end retaining ring.

[0042] The clearance between the outer spring 7 and the outer rod 1 is increased to 0.2-0.4 mm, allowing the spring to produce moderate yaw during vertical vibration; the inner spring 8 and the inner rod 2 are interference-fitted with an interference of 0.05-0.08 mm to ensure horizontal constraint rigidity. The trapezoidal inclined surface of the outer rod protrusion 3 forms a wedge-shaped contact with the spring coil, generating a self-centering effect when the spring is subjected to lateral force.

[0043] In this embodiment, the outer rod protrusion 3 of the helical array periodically moves along the helical path of contact with the spring under the condition of variable compressor speed. When the core speed is 2000 r / min, the collision frequency between the protrusion and the spring is set to 33Hz (i.e., 3-line helix × 2000 / 60), which is far from the first natural frequency (45-50Hz) of common compressor housings, thus avoiding resonance amplification. The inner rod protrusion 4 of the linear array generates a uniform constraint force with a fixed pitch, providing stable horizontal limit in a wide frequency range of 0-500Hz. The inclined surface of the helical protrusion generates direction-dependent damping for the spring movement. When the core is deflected to the right by horizontal centrifugal force, the inclined surface of the right-hand helical protrusion causes the spring to rebound to the left; conversely, when deflected to the left, the inclined surface of the protrusion generates a rightward restoring force. This asymmetric damping characteristic increases the horizontal vibration decay rate by 25%, especially during the compressor start-up and shutdown phase, reducing the maximum core offset from 1.2mm to 0.7mm. The inner rod protrusion 4 of the linear array has a semi-circular cross-section. When the temperature rises and the nylon bushing expands, the protrusion radius increases by 0.02-0.03 mm, which increases synchronously with the interference fit of the inner spring 8, compensating for the loss of preload caused by thermal expansion. The spacing of the helical protrusions on the outer rod 1 is slightly adjusted with temperature changes. When the spacing increases by 0.1%, the helix angle automatically decreases by 0.3° to maintain the stability of the contact pressure with the spring.

[0044] Compared to Example 1, the trapezoidal bevel of the spiral protrusion in this example disperses the spring contact stress over a larger area, reducing the maximum contact pressure from 350 MPa to 220 MPa, which is lower than the yield strength of nylon material (280 MPa). The semi-circular cross-section of the straight protrusion reduces the stress concentration factor from 2.1 to 1.4, extending the fatigue life to 10. ^7 After several cycles, under abnormal operating conditions where the compressor instantaneously overspeeds to 5000 r / min, the multi-contact point distribution of the spiral protrusions reduces the single-point impact load by 40%, and combined with the uniformly distributed constraint of the straight protrusions, effectively prevents spring instability. Tests show that this structure can still maintain a vibration suppression efficiency of over 85% at a low temperature of -30℃.

[0045] This embodiment achieves targeted optimization of vibration control performance at the same material cost through an innovative raised distribution structure. Bench tests have verified that compressors using this bushing exhibit a 4-6 dB reduction in sound pressure level in the 200-800 Hz frequency band during variable frequency operation, and a 30% reduction in axial vibration transmissibility, making it particularly suitable for noise-sensitive environments.

[0046] Example 3

[0047] This embodiment further optimizes the structural dimensions and stress distribution characteristics based on the core architecture of the double-layer bushing 100 described in Embodiment 1. The outer rod 1 and the inner rod 2 of the spring bushing adopt a differentiated diameter design. The outer circumferential diameter of the outer rod 1 is increased by a specific proportion compared with that of the inner rod 2. At the same time, the bushing wall thickness varies along the axial direction, forming a gradient structure that is thicker at the bottom and thinner at the top.

[0048] The outer diameter of the outer rod 1 is set to be 1.8 times the diameter of the inner rod 2, and the ratio is determined based on the difference in wire diameter between the outer spring 7 and the inner spring 8.

[0049] Specifically, in this embodiment, the outer spring wire diameter is 1.5mm, the inner spring wire diameter is 1.0mm, and the difference between the two is 0.5mm. The difference in the outer rod diameter is 0.5mm × 1.8 = 0.9mm. This design ensures that the helical gaps of the inner and outer springs do not interfere with each other when the two types of springs are nested together, while maintaining sufficient radial space for spring deflection. The diameter expansion area of ​​the outer rod 1 is concentrated in the lower part of the bushing, gradually narrowing at the top to smoothly transition with the inner rod 2, forming a streamlined profile and reducing airflow disturbance noise.

[0050] The wall thickness of the double-layer bushing 100 decreases from bottom to top: 3.0 mm at the bottom, 2.2 mm in the middle transition area, and 1.8 mm at the top. The height of each area is approximately one-third of the total bushing length, with a 2.0 mm radius rounded transition at the junctions. The thickened bottom design provides higher compressive strength to the interference fit area between the bushing and the protruding stud 6, increasing the measured compressive load from 1200 N to 1800 N. The thinned top reduces the overall weight, making the bushing 15% lighter than in Example 1, while maintaining sufficient axial stiffness. The wall thickness gradient is achieved through a slider mechanism in the injection mold, ensuring that the internal cavity dimensional accuracy is controlled within ±0.05 mm after molding.

[0051] During compressor operation, the diameter difference between the outer rod 1 and the inner rod 2 generates a synergistic vibration reduction effect. The outer spring 7 gains greater radial movement space due to the increased diameter of the outer rod, improving vertical vibration absorption efficiency by 20%. Meanwhile, the smaller diameter of the inner rod 2 concentrates the preload of the inner spring 8 in the horizontal direction, limiting the lateral displacement of the compressor core to within ±0.3mm. When the compressor speed jumps, the gradually varying wall thickness structure adjusts the vibration transmission path through differentiated local stiffness: the thick-walled area at the bottom suppresses low-frequency vibration transmission, the transition area in the middle attenuates mid-frequency resonance, and the thin-walled area at the top weakens high-frequency noise radiation.

[0052] During actual assembly, the clearance between the outer rod 1 and the outer spring 7 is adjusted to 0.3-0.5mm to accommodate dynamic deformation after the diameter increases. Operators use a special flaring tool to widen the spring end to 1.05 times the diameter of the outer rod before inserting it, utilizing the spring's rebound force to create a uniform gap. The inner spring 8 is installed using a hot-pressing process: the inner rod 2 is heated to 80℃ to expand by 0.1mm, and after cooling and shrinking, a stable interference fit is formed. The gradual wall thickness design shortens the injection molding cycle by 12% and increases the demolding yield to over 99%.

[0053] Vibration table testing demonstrated that the double-layer bushing exhibited excellent stability across the entire compressor speed range. The peak vertical vibration acceleration decreased from 2.8 m / s² to 0.9 m / s², and the horizontal vibration energy decreased by 55%. Noise spectrum analysis showed an 8 dB(A) reduction in high-frequency components above 500 Hz, bringing the overall sound quality close to the requirements of a silent refrigerator. Long-term durability testing showed that the gradually increasing wall thickness effectively dispersed alternating stress, and the bushing showed no visible cracks after 100,000 cycles, achieving a lifespan 1.5 times that of the industry standard.

[0054] This embodiment achieves a comprehensive improvement in vibration control performance within a limited space through precise matching of structural dimensions and wall thickness. It is particularly suitable for scenarios with extremely high requirements for space utilization and reliability, such as vehicle refrigerator compressors.

Claims

1. A spring bushing for a small refrigeration compressor, characterized in that: It includes a double-layer bushing (100), the inner cavity of which is adapted to the protrusion (6) of the lower housing (5) of the compressor to form an interference fit; the double-layer bushing (100) includes an outer rod (1) and an inner rod (2) arranged coaxially, the outer rod (1) is covered with an outer spring (7), and the inner rod (2) is covered with an inner spring (8); the stiffness of the inner spring (8) is greater than that of the outer spring (7).

2. The spring bushing for a small refrigeration compressor according to claim 1, characterized in that: The outer rod body (1) has circumferentially distributed outer rod protrusions (3) on its outer circumferential surface, and the inner rod body (2) has circumferentially distributed inner rod protrusions (4) on its outer circumferential surface. The top heights of the outer rod protrusions (3) and the inner rod protrusions (4) respectively extend beyond the outer circumferential surface of the corresponding rod body.

3. The spring bushing for a small refrigeration compressor according to claim 2, characterized in that: The cross-sectional shapes of the outer rod protrusion (3) and the inner rod protrusion (4) are arc-shaped or trapezoidal, and their shapes may be the same or different.

4. The spring bushing for a small refrigeration compressor according to claim 1, characterized in that: The outer rod (1) and the outer spring (7), and the inner rod (2) and the inner spring (8) are respectively fitted with clearance or interference fit.

5. The spring bushing for a small refrigeration compressor according to claim 1, characterized in that: The bottom end face of the double-layer bushing (100) forms a plane contact with the mounting base surface of the protruding nail (6), and the connection between the outer rod body (1) and the inner rod body (2) is provided with annular reinforcing ribs.

6. The spring bushing for a small refrigeration compressor according to claim 2, characterized in that: The outer rod protrusions (3) are arranged in a spiral array along the axial direction of the outer rod body (1), and the inner rod protrusions (4) are arranged in a straight array with equal spacing along the axial direction of the inner rod body (2).

7. The spring bushing for a small refrigeration compressor according to any one of claims 1-6, characterized in that: The outer diameter of the outer rod (1) is greater than the outer diameter of the inner rod (2), and the difference between the two diameters is 1.1-3 times the difference between the wire diameters of the outer spring (7) and the inner spring (8).

8. The spring bushing for a small refrigeration compressor according to any one of claims 1-6, characterized in that: The wall thickness of the double-layer bushing (100) gradually decreases from bottom to top, and the bottom wall thickness is 1.1-3 times the top wall thickness.

9. The spring bushing for a small refrigeration compressor according to any one of claims 1-6, characterized in that: The outer rod (1) and inner rod (2) of the double-layer bushing (100) are integrally formed.

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