A compressor

Abstract

The present invention relates to a compressor, in particular to a rolling rotor type compressor; the compressor includes a crankshaft rotor assembly, and the crankshaft rotor assembly includes a rotor and a crankshaft; the present application scheme determines the factors affecting the natural frequency of the crankshaft rotor assembly through experimental verification, sets the optimal frequency range of the natural frequency of the crankshaft rotor assembly, and proposes specific parameter settings of the crankshaft to reduce the low-frequency vibration of the compressor, thereby solving the low-frequency noise of the existing compressor and reducing the low-frequency noise generated when the compressor is running.

Description

Technical Field

[0001] The present invention relates to the technical field of compressors, and in particular to a compressor. Background Art

[0002] Rolling rotor variable frequency compressors have outstanding advantages such as small size and simple structure, and are widely used in household air conditioners, commercial air conditioners, low-temperature heat pumps and other fields. In the field of air conditioning and refrigeration, consumers have higher and higher requirements for outdoor unit noise in addition to cooling and heating. The compressor is the main noise source of the outdoor unit. By wrapping the outside of the compressor with sound insulation cotton, adding a silencer inside the compressor, and adding the sound insulation effect of the wall, medium and high frequency noise is difficult to enter the room. However, the peak low-frequency noise of the compressor has strong penetrating power and can easily enter the room with unpleasant sound quality, which has become a major complaint of consumers. The current industry's common practice to solve this problem is to shield the operating frequency with high low-frequency noise peak, but sometimes when too many operating frequency points are shielded, the air conditioner's capacity adjustment continuity will deteriorate and the room temperature fluctuation will increase.

[0003] The low-frequency noise of the compressor is usually within 400Hz on the sound spectrum; when the compressor is running, the fluctuation of gas force, exhaust pulsation and other factors cause the compressor to vibrate and transmit to the outdoor unit pipe and casing; when the compressor runs at certain frequency points, it excites certain parts of the compressor to resonate and causes the outdoor unit pipe or casing to resonate, and at this time the peak value of the low-frequency noise of the outdoor unit will increase significantly; through statistical analysis, it is found that the frequency of these low-frequency noises is usually 2 times, 3 times, 4 times, etc. of the compressor operating frequency. Summary of the Invention

[0004] The technical problem to be solved by the present invention is to overcome the deficiencies of the prior art and provide a compressor, aiming to reduce the low-frequency noise of the existing compressor.

[0005] The present invention aims to design a compressor, which includes a crankshaft rotor assembly, wherein the crankshaft rotor assembly includes a rotor and a crankshaft; the natural frequency F of the crankshaft rotor assembly is qz Designed according to the following formula:

[0006] λ is a conventional coefficient, D is the shaft diameter of the crankshaft 7 , and m is the mass of the crankshaft rotor assembly 3 .

[0007] In some embodiments, when the compressor is a single-cylinder compressor, the natural frequency F qz satisfy:

[0008] 20+F max <F qz <2(F max -10) or F qz >2(F max +10), F maxThe maximum operating frequency of the compressor.

[0009] In some embodiments, when the compressor maximum operating frequency F max When the frequency is 120Hz, F qz Satisfy: 140Hz<F qz <220Hz or F qz >260Hz.

[0010] In some embodiments, the compressor is a two-cylinder compressor; when the total displacement of the compressor is less than 36cc, the natural frequency F qz Satisfied: F qz <2(F max -10); When the total displacement of the compressor is greater than 36cc, the natural frequency F qz Satisfaction: 2(F max +10)<F qz <3(F max -10).

[0011] In some embodiments, the compressor is a three-cylinder compressor; when the compressor has a maximum operating frequency F max When ≤90Hz, the natural frequency F qz Satisfied: F qz >3(F max +10); when the compressor's maximum operating frequency F max When >90Hz, the natural frequency F qz Satisfaction: 2(F max +10)<F qz <3(F max -10).

[0012] In some embodiments, the diameter D of the crankshaft 7 and the mass m of the crankshaft rotor assembly 3 have a value range of:

[0013] 10mm≤D≤14mm, 0.4kg≤m≤0.85, or 14 <D,m≤0.4kg。

[0014] In some embodiments, when the total displacement of the compressor is less than 36cc, the range of the diameter D of the crankshaft 7 and the mass m of the crankshaft rotor assembly 3 is: D ≤ 16mm, m > 0.7kg; when the total displacement of the compressor is greater than 36cc, the range of the diameter D of the crankshaft 7 and the mass m of the crankshaft rotor assembly 3 is: 17mm <D≤24mm,0.7kg<m<2.5kg。

[0015] In some embodiments, when the compressor maximum operating frequency F max ≤90Hz, the range of the diameter D of the crankshaft 7 and the mass m of the crankshaft rotor assembly 3 is: 23mm≤D, m≥1.6kg; when the maximum operating frequency of the compressor is Fmax When the frequency is higher than 90 Hz, the range of the diameter D of the crankshaft 7 and the mass m of the crankshaft rotor assembly 3 is: 17 mm ≤ D < 23 mm, 0.8 kg <m<2.6kg。

[0016] In some embodiments, an annular groove is provided on the crankshaft, and the annular groove is located between the rotor assembly 3 and the upper end surface of the upper flange; the compressor adjusts the natural frequency F of the crankshaft rotor assembly by the diameter value d of the annular groove. qz ; Natural frequency F qz satisfy:

[0017] In some embodiments, the compressor is a rolling rotor compressor; the value range of λ is: 0.75-0.85.

[0018] The present invention proposes a compressor, which verifies and determines the factors affecting the natural frequency of the crankshaft rotor assembly through experiments, sets the optimal frequency range of the natural frequency of the crankshaft rotor assembly, and proposes specific parameter settings of the crankshaft to reduce the low-frequency vibration of the compressor, thereby solving the low-frequency noise of the existing compressor and reducing the low-frequency noise generated during the operation of the compressor.

[0019] It is to be understood that the foregoing general description and the following detailed description are exemplary and explanatory only and are not restrictive of the invention. BRIEF DESCRIPTION OF THE DRAWINGS

[0020] The accompanying drawings are part of the present invention and are used to provide a further understanding of the present invention. The exemplary embodiments of the present invention and their descriptions are used to explain the present invention, but do not constitute an undue limitation of the present invention. Obviously, the drawings described below are only some embodiments. For those of ordinary skill in the art, other drawings can be obtained based on these drawings without inventive effort. In the accompanying drawings:

[0021] Figure 1 is a schematic structural diagram of a single-cylinder compressor according to an exemplary embodiment of the present invention;

[0022] Figure 2 is a schematic structural diagram of a twin-cylinder compressor according to an exemplary embodiment of the present invention;

[0023] Figure 3 is a schematic structural diagram of a three-cylinder compressor according to an exemplary embodiment of the present invention;

[0024] Figure 4 is a schematic diagram of a crankshaft structure according to an exemplary embodiment of the present invention;

[0025] Figure 5 is a gas torque fluctuation characteristic diagram of a single-cylinder compressor according to an exemplary embodiment of the present invention;

[0026] Figure 6 is a gas torque fluctuation characteristic diagram of a twin-cylinder compressor according to an exemplary embodiment of the present invention;

[0027] Figure 7 is a gas torque fluctuation characteristic diagram of a three-cylinder compressor according to an exemplary embodiment of the present invention;

[0028] Figure 8 3f is a relationship between the magnitudes of the excitation forces before the single-cylinder compressor according to an exemplary embodiment of the present invention;

[0029] Figure 9 3f is the magnitude relationship of the excitation force of the front 3f of a twin-cylinder compressor according to an exemplary embodiment of the present invention;

[0030] Figure 10 3f is the magnitude relationship of the excitation force before the three-cylinder compressor according to an exemplary embodiment of the present invention;

[0031] Figure 11 The present invention is an exemplary embodiment of improving the housing vibration characteristics of a front single-cylinder compressor crankshaft rotor assembly when it is excited to resonate by a 1f excitation force;

[0032] Figure 12 Schematic diagram showing that after improvement according to an exemplary embodiment of the present invention, the natural frequency of the crankshaft rotor assembly is increased and resonance is not excited by the 1f excitation force within Fmax;

[0033] Figure 13 1 is a schematic diagram showing a first solution for increasing the natural frequency of a crankshaft rotor assembly in a single-cylinder compressor according to an exemplary embodiment of the present invention and its effect;

[0034] Figure 14 1 is a schematic diagram showing a second solution for increasing the natural frequency of a crankshaft rotor assembly in a single-cylinder compressor according to an exemplary embodiment of the present invention and its effect;

[0035] Figure 15 1 is a schematic diagram showing the setting of the natural frequency of the crankshaft rotor assembly when the displacement of a two-cylinder compressor is less than 36cc according to an exemplary embodiment of the present invention;

[0036] Figure 16 1 is a schematic diagram showing the setting of the natural frequency of the crankshaft rotor assembly when the displacement of a two-cylinder compressor is greater than 36cc according to an exemplary embodiment of the present invention;

[0037] Figure 17 2 is a schematic diagram showing the setting of the natural frequency of the crankshaft rotor assembly when the maximum operating frequency of the three-cylinder compressor is less than 90 Hz according to an exemplary embodiment of the present invention;

[0038] Figure 183 is a schematic diagram showing the setting of the natural frequency of the crankshaft rotor assembly when the maximum operating frequency of the three-cylinder compressor is greater than 90 Hz according to an exemplary embodiment of the present invention.

[0039] In the figure: 1. Housing; 2. Motor; 3. Crankshaft-rotor assembly; 4. Pump body assembly; 5. Annular stator; 6. Rotor; 7. Crankshaft; 71. Annular groove; 81. Upper flange; 82. Lower flange; 9. Cylinder; 91. Upper cylinder; 92. Lower cylinder; 93. Middle cylinder; 10. Liquid distributor; 11. Partition; 12. Upper partition; 13. Middle partition; 14. Lower partition.

[0040] It should be noted that these drawings and textual descriptions are not intended to limit the conceptual scope of the present invention in any way, but rather to illustrate the concept of the present invention for those skilled in the art by referring to specific embodiments. DETAILED DESCRIPTION

[0041] In the description of the present invention, it should be noted that the terms "inside" and "outside" etc. indicating orientations or positional relationships are based on the orientations or positional relationships shown in the accompanying drawings, and are only for the convenience of describing the present invention and simplifying the description, rather than indicating or implying that the device or element referred to must have a specific orientation, be constructed and operated in a specific orientation, and therefore cannot be understood as a limitation on the present invention.

[0042] In the description of the present invention, it should be noted that, unless otherwise expressly specified or limited, the terms "mounted," "connected," "connected," "in contact," and "connected" should be understood broadly. For example, they may refer to fixed, detachable, or integral connections; mechanical or electrical connections; and direct or indirect connections through an intermediary. Those skilled in the art will understand the specific meanings of the above terms in the present invention based on the specific circumstances.

[0043] The present invention provides a compressor, in particular a rolling rotor type compressor, which solves the problem of low-frequency noise during operation of existing compressors.

[0044] like Figures 1-4As shown, the present invention proposes a compressor, specifically a rolling rotor variable frequency compressor; the compressor includes a liquid separator 10, a housing 1, a motor 2, and a pump body assembly 4; the liquid separator 10 is arranged on the outside of the housing 1, and is used to transport low-pressure refrigerant to the pump body assembly 4. The upper part of the liquid separator 10 is fixed to the housing 1 through a liquid separator bracket, and the lower part passes through the housing 1 through a bend to connect to the pump body suction port; the motor 2 and the pump body assembly 4 are wrapped in the housing 1 and form an airtight isolation with the external environment; specifically, the motor 2 includes an annular stator 5 and a rotor 6, the annular stator 5 is fixed to the upper part of the housing 1 with an interference fit, and the pump body assembly 4 It is arranged at the lower part of the annular stator 5; further, the pump body assembly 5 includes an upper flange 81, a cylinder 9, a partition 11, a lower flange 82, a crankshaft 7, a roller and a vane; the rotor 6 is arranged in the annular stator 5 and is sleeved on the crankshaft 7 with an interference fit, and the motor 2 drives the crankshaft 7 through the rotor 6 to operate so that the pump body assembly 4 sucks the low-pressure refrigerant in the liquid distributor 10 and discharges it from the pump body after compression, and the output range of the compressor capacity can be frequency-controlled; further, the crankshaft rotor assembly 3 includes a rotor 6 and a crankshaft 7. The resonance of the crankshaft rotor assembly 3 is an important factor causing the large low-frequency noise of the compressor in the present application; as Figure 1 As shown, the single-cylinder compressor includes a cylinder 9, which is clamped and fixed by an upper flange 81 and a lower flange 82 respectively; Figure 2 As shown, the twin-cylinder compressor includes an upper cylinder 91 and a lower cylinder 92, and the upper cylinder 91 and the lower cylinder 92 are separated by a partition 11; Figure 3 As shown, the three-cylinder compressor includes an upper cylinder 91, a lower cylinder 92 and a middle cylinder 93. An upper partition 12 and a middle partition 13 are provided between the upper cylinder 91 and the middle cylinder 93, and a lower partition 14 is provided between the middle cylinder 93 and the lower cylinder 92.

[0045] Specifically, if Figures 1-4As shown in FIG, the influence of the natural frequency of the crankshaft rotor assembly of compressors with different numbers of cylinders on the vibration of the compressor: the natural frequency of the crankshaft rotor assembly is its inherent property. The crankshaft rotor assemblies with different structures all have natural frequencies. The main difference is that the natural frequencies are large or small. When the frequency of the excitation force is close to the natural frequency of the crankshaft rotor assembly, the crankshaft rotor assembly will be stimulated to resonate, which will cause a significant increase in the vibration acceleration and noise near the frequency point. For rotating machinery such as the rolling rotor compressor of the present invention, the frequency of the excitation force generated during operation is usually an integer multiple of the operating frequency f. If the operating frequency of the compressor is f, 1 f, 2f, 3f, 4f, 5f, ..., nf frequency-multiplied excitation forces. The generation principles of different frequency-multiplied excitation forces of different compressor structures are different, and the corresponding excitation forces will also be different. The frequencies of different frequency-multiplied excitation forces are close to the natural frequency of the crankshaft rotor and cause the vibration acceleration values ​​generated after the crankshaft rotor assembly resonates. For example, for a rotor compressor with only one cylinder (single-cylinder compressor), its 1f excitation force is mainly caused by the torque fluctuation and exhaust pulsation of the cylinder compressed gas, and the corresponding 1f excitation force is relatively large; the 2f excitation force is mainly generated by the harmonics of the 1f excitation force, and the corresponding 2f excitation force is relatively small (such as Figure 5 、 Figure 8 ); For a rotor compressor with two cylinders (twin-cylinder compressor), its 1f excitation force is mainly caused by the centrifugal force imbalance generated by the crankshaft, roller eccentric mass, and the eccentric mass of the main and auxiliary balance blocks on the rotor when the crankshaft rotates. Its 1f excitation force is smaller than that of a single-cylinder compressor; while the 2f excitation force is mainly caused by the torque fluctuation and exhaust pulsation of the compressed gas in the upper and lower cylinders. The corresponding 2-harmonic frequency excitation force is larger than that of a single-cylinder compressor ( Figure 6 、 Figure 9 ); For a rotor compressor with three cylinders (three-cylinder compressor), its 1f excitation force is mainly caused by the centrifugal force imbalance generated by the eccentric mass of the crankshaft, roller, and the eccentric mass of the main and auxiliary balance blocks on the rotor when the crankshaft rotates. Its 1f excitation force is smaller than that of a single-cylinder compressor; the 2f excitation force is mainly generated by the harmonics of the 1f excitation force, and its 2f excitation force is smaller than that of a two-cylinder compressor. The 3f excitation force is mainly caused by the torque fluctuation and exhaust pulsation of the compressed gas in the upper, middle and lower cylinders, so its 3f excitation force is larger than the 3f excitation force of both single-cylinder and two-cylinder compressors (such as Figure 7 、 Figure 10); If both the 1f excitation force and the 2f excitation force cause resonance in the crankshaft rotor assembly of a single-cylinder compressor, the vibration acceleration value of the compressor caused by the resonance of the 1f excitation force is greater than the vibration acceleration value of the compressor caused by the 2f excitation force; If both the 1f excitation force and the 2f excitation force cause resonance in the crankshaft rotor assembly of a two-cylinder compressor, the vibration acceleration value of the compressor caused by the resonance of the 1f excitation force is less than the vibration acceleration value of the compressor caused by the 2f excitation force; If the 1f, 2f, and 3f excitation forces of a three-cylinder compressor all cause resonance in the crankshaft rotor assembly, the vibration acceleration generated by the resonance of the crankshaft rotor assembly caused by the 3f excitation force is greater than the vibration acceleration caused by the 1f and 2f excitation forces. Therefore, the natural frequency value of the crankshaft rotor assembly should be designed specifically according to the number of cylinders in the compressor.

[0046] Preferably, in combination with the above scheme, Figures 1-4 As shown, the compressor includes a crankshaft rotor assembly 3, wherein the crankshaft rotor assembly 3 includes a rotor 6 and a crankshaft 7; specifically, the natural frequency F of the crankshaft rotor assembly 3 is qz Designed according to the following formula:

[0047] Where: λ is the conventional coefficient, which can be 0.75-0.85; D is the shaft diameter of the crankshaft 7, and m is the mass of the crankshaft rotor assembly 3. Specifically, the conventional coefficient λ can be designed and rotated according to the actual crankshaft. For example, if the shaft diameter D of the crankshaft major axis is designed to be 19 mm and the mass of the rotor assembly is 1.4 kg, then its fixed frequency is approximately 238 Hz, and the value of λ is 0.78.

[0048] By adopting the above scheme, the natural frequency of the crankshaft rotor assembly is designed It can effectively reduce the low-frequency noise when the compressor is running.

[0049] Preferably, in combination with the above scheme, Figure 1 As shown, when the compressor is a single-cylinder compressor, the natural frequency F qz satisfy:

[0050] 20+F max <F qz <2(F max -10) or F qz >2(F max +10); where: F max The maximum operating frequency of the compressor.

[0051] Preferably, in combination with the above scheme, Figure 1 As shown, when the compressor maximum operating frequency F max When the frequency is 120Hz, F qz Satisfy: 140Hz<F qz <220Hz or F qz >260Hz.

[0052] Preferably, in combination with the above scheme, Figure 2 As shown, the compressor is a two-cylinder compressor; when the total displacement of the compressor is less than 36cc, the natural frequency F qz Satisfied: F qz <2(F max -10); When the total displacement of the compressor is greater than 36cc, the natural frequency F qz Satisfaction: 2(F max +10)<F qz <3(F max -10).

[0053] Preferably, in combination with the above scheme, Figure 3 As shown, the compressor is a three-cylinder compressor; when the compressor has the maximum operating frequency F max When ≤90Hz, the natural frequency F qz Satisfied: F qz >3(F max +10); when the compressor's maximum operating frequency F max When >90Hz, the natural frequency F qz Satisfaction: 2(F max +10)<F qz <3(F max -10).

[0054] Preferably, in combination with the above scheme, Figure 1 As shown in the figure, when the compressor is a single-cylinder compressor, the natural frequency is: The range of the diameter D of the crankshaft 7 and the mass m of the crankshaft rotor assembly 3 is: 10mm≤D≤14mm, 0.4kg≤m≤0.85, or 14 <D,m≤0.4kg。

[0055] Preferably, in combination with the above scheme, Figure 2 As shown in the figure, when the compressor is a two-cylinder compressor, the natural frequency is: When the total displacement of the compressor is less than 36cc, the range of the diameter D of the crankshaft 7 and the mass m of the crankshaft rotor assembly 3 is: D≤16mm, m>0.7kg; further, when the total displacement of the compressor is greater than 36cc, the range of the diameter D of the crankshaft 7 and the mass m of the crankshaft rotor assembly 3 is: 17mm <D≤24mm,0.7kg<m<2.5kg。

[0056] Preferably, in combination with the above scheme, Figure 3 As shown in the figure, when the compressor is a two-cylinder compressor, the natural frequency is: When the compressor's maximum operating frequency F max≤90Hz, the range of the diameter D of the crankshaft 7 and the mass m of the crankshaft rotor assembly 3 is: 23mm≤D, m≥1.6kg; further, when the maximum operating frequency of the compressor F max When the frequency is higher than 90 Hz, the range of the diameter D of the crankshaft 7 and the mass m of the crankshaft rotor assembly 3 is: 17 mm ≤ D < 23 mm, 0.8 kg <m<2.6kg。

[0057] Preferably, in combination with the above scheme, Figure 4 As shown, in this embodiment, an annular groove 71 is set at a specific position of the crankshaft 7 to adjust the natural frequency of the crankshaft rotor assembly to an appropriate value, that is, the crankshaft 7 is provided with an annular groove 71, and the annular groove 71 is located between the rotor assembly 3 and the upper end surface of the upper flange; specifically, the compressor adjusts the natural frequency F of the crankshaft rotor assembly by the diameter value d of the annular groove 71 (that is, the crankshaft diameter within the annular groove 71). qz ; Natural frequency F qz satisfy: Specifically:

[0058] When the compressor is a single-cylinder compressor, adjust the annular groove diameter value d (i.e. the crankshaft diameter in the annular groove) to meet 20+F max <F qz <2(F max -10) or F qz >2(F max +10); when the maximum operating frequency of the compressor is 120Hz, F qz Satisfy: 140Hz<F qz <220Hz or F qz >260Hz;

[0059] When the compressor is a twin-cylinder compressor, when the total displacement is less than 36cc, F qz Should meet: F qz <2(F max -10); when the total displacement is greater than 36cc, F qz Should meet: 2(F max +10)<F qz <3(F max -10);

[0060] When the compressor is a three-cylinder compressor, when F max ≤90Hz, then F qz Should meet: F qz >3(F max +10); when F max >90Hz, then F qz Should meet: 2(F max +10)<F qz <3(F max-10).

[0061] Specifically, in combination with the above solution, Figure 1 、 Figure 5 、 Figure 8 As shown, the natural frequency F of the crankshaft rotor assembly of a single-cylinder compressor max Value range: In a single-cylinder compressor, the 1f excitation force is the largest, followed by the 2f excitation force, so the natural frequency of the crankshaft rotor assembly F qz The value must be greater than the maximum operating frequency F of the compressor max (Usually the maximum operating frequency of the rotor compressor is 120Hz). When the difference between the natural frequency of the crankshaft rotor assembly and the operating frequency of the compressor is less than 40Hz, the vibration acceleration of the compressor 1f will increase significantly. Therefore, in order to avoid the above problem, the natural frequency of the crankshaft rotor assembly F max Should satisfy the formula: F qz >20+F max (like Figure 12 ); Furthermore, when the vibration acceleration of the crankshaft rotor assembly caused by the 2f excitation force is less than the vibration acceleration caused by the 1f excitation force, but the natural frequency value of the crankshaft rotor assembly is not set properly, the compressor also needs to shield some frequency points when running on the air conditioner, which may lead to insufficient maximum capacity output of the compressor; for example, the following situation occurs: the operating frequency of the crankshaft rotor assembly excited by the 2f excitation force of the compressor is just at the maximum operating frequency F max If these operating frequency points are blocked, the maximum output of the compressor will be reduced, affecting the cooling and heating effects of the air conditioner. If these operating frequency points are not blocked, the 2f vibration acceleration and noise peak of the compressor will be high when running at these frequency points, affecting the comfort of consumers. This application adopts two solutions:

[0062] like Figure 13 The first method is to appropriately reduce the natural frequency of the crankshaft rotor assembly so that the 2f excitation force does not excite the resonance of the crankshaft rotor assembly when the compressor is near the highest operating frequency. When the vibration acceleration at the operating frequency point that causes the crankshaft assembly to produce 2f resonance is large, some frequencies are appropriately shielded so that the maximum output capacity of the compressor is not affected. To achieve this effect, the natural frequency of the crankshaft rotor assembly F qz Also need to meet: F qz <2(F max -10);

[0063] like Figure 14 The second method is to increase the natural frequency of the crankshaft rotor assembly. Even if the compressor is running at the highest operating frequency, the 2f excitation force will not excite the resonance of the crankshaft rotor assembly. To achieve this effect, the natural frequency of the crankshaft rotor assembly F qz Also need to meet: F qz >2(F max +10);

[0064] Therefore, for a single-cylinder compressor, the natural frequency of the crankshaft rotor assembly is qz Should meet: 20+F max <F qz <2(F max -10) or F qz >2(F max +10), the maximum operating frequency of the compressor is usually 120Hz. For a single-cylinder compressor, the natural frequency of the crankshaft rotor assembly is F qz If set to: (140Hz, 220Hz) or F qz When the frequency is higher than 260Hz, the low-frequency vibration and noise of the compressor caused by the resonance of the crankshaft rotor assembly can be greatly reduced, achieving a better noise reduction effect.

[0065] Specifically, in combination with the above solution, Figure 2 、 Figure 6 、 Figure 9 As shown, the natural frequency F of the crankshaft rotor assembly of the two-cylinder compressor qz Value range: The 1f excitation force of a two-cylinder compressor is relatively small (which can be improved by optimizing the balance block on the rotor assembly), the 2f excitation force is the largest, and the 3f excitation force is the second largest. Therefore, in order to avoid the compressor from exciting the resonance of the crankshaft rotor assembly within the operating frequency range, the ideal way to reduce the resonance of the crankshaft rotor assembly caused by the 2f excitation force is to make the natural frequency of the crankshaft rotor assembly greater than 2 times the maximum operating frequency of the compressor (i.e. F qz >2F max ); by the formula It can be seen that the most important factor affecting the natural frequency of the crankshaft rotor assembly is the crankshaft diameter. The larger the diameter, the higher the natural frequency F. qz The higher the value, the larger the crankshaft diameter should be. However, the larger the crankshaft diameter is, the lower the compressor performance will be. Therefore, the crankshaft diameter should not be too large. Considering the effects of improving performance and reducing the resonance of the crankshaft rotor assembly, it is easy to achieve for compressors with a total displacement of more than 36cc. However, it is difficult to meet the requirements for twin-cylinder compressors with a total displacement of less than 36cc. Therefore, in order to improve performance and reduce the crankshaft diameter, the natural frequency F of the crankshaft rotor assembly should be reduced. qz Reduce and be within the frequency range of resonance excited by the compressor 2f excitation force; In order to reduce the influence of the resonance of the crankshaft rotor assembly, increasing the crankshaft diameter will easily lead to reduced compressor performance; Therefore, it is necessary to carry out targeted design according to the difference in compressor displacement.

[0066] Specifically, if Figure 15When the total displacement of the compressor is less than 36cc, in order to avoid the compressor exciting the resonance of the crankshaft rotor assembly near the maximum operating frequency and shielding the frequency range, thereby affecting the output of the maximum capacity of the compressor, the natural frequency of the crankshaft rotor assembly F qz The appropriate frequency range should be F qz <2(F max -10). At this time, although some frequency points may need to be shielded due to the resonance of the crankshaft rotor assembly within the maximum operating frequency range of the compressor, the maximum operating frequency point is avoided from being shielded, and the maximum output capacity of the compressor is not affected, and a good overall effect can be achieved.

[0067] Specifically, if Figure 16 When the total displacement of the compressor is greater than 36cc, the F can be satisfied by appropriately increasing the crankshaft diameter. qz >2(F max +10). However, an excessively large crankshaft diameter may cause two problems: first, it may cause a decrease in compressor performance, which is inconsistent with the trend and requirements of energy conservation and emission reduction; second, the natural frequency F of the crankshaft rotor assembly may increase. qz If the F is too large, it will also cause the compressor 3f excitation force to excite the crankshaft rotor assembly resonance at a higher operating frequency. If the operating frequency at this time is close to the maximum operating frequency, it needs to be shielded and the maximum capacity output will be affected. Therefore, F qz Should meet: F qz <3(F max -10);

[0068] In summary, for a twin-cylinder compressor:

[0069] When the total displacement of the compressor is less than 36cc, F qz Should meet: F qz <2(F max -10);

[0070] When the total displacement of the compressor is greater than 36cc, F qz Should meet: 2(F max +10)<F qz <3(F max -10).

[0071] Specifically, in combination with the above solution, Figure 3 、 Figure 7 、 Figure 10 As shown, the natural frequency F of the crankshaft rotor assembly of the three-cylinder compressor qzValue Range: The maximum excitation force during operation of a three-cylinder compressor is the 3f excitation force. Therefore, the natural frequency of the crankshaft rotor assembly should be set primarily to reduce the impact of the 3f excitation force on its resonance. When the maximum operating frequency of the compressor is low, such as at 90 Hz, the 3f excitation force is 270 Hz. In this case, the natural frequency of the crankshaft rotor assembly can be increased above the 3f excitation force by appropriately increasing the shaft diameter or reducing the mass of the rotor assembly, thereby causing the 3f excitation force to excite the crankshaft rotor assembly to resonate. However, when the maximum operating frequency of the compressor is high, such as above 90 Hz, the natural frequency of the crankshaft rotor assembly needs to be higher to avoid resonance of the crankshaft rotor assembly caused by the 3f excitation force when operating above 90 Hz. In this case, increasing the crankshaft diameter or reducing the rotor assembly mass will result in reduced compressor performance or require an increase in the compressor casing diameter due to the increased crankshaft diameter, thereby increasing the compressor cost. By appropriately reducing the natural frequency of the crankshaft rotor assembly and adjusting the operating frequency corresponding to the resonance of the crankshaft rotor assembly to an uncommon frequency point, the maximum output capacity of the compressor will not be lost, and there will be no reduction in compressor performance or increase in compressor cost.

[0072] Specifically, if Figure 16 When the maximum operating frequency of the compressor is less than or equal to 90Hz, the natural frequency of the crankshaft rotor assembly F can be set qz Satisfied: F qz >3(F max +10);

[0073] Specifically, if Figure 17 When the maximum operating frequency of the compressor is greater than 90Hz, the natural frequency of the crankshaft rotor assembly F can be set. qz Satisfied: F qz <3(F max -10), and in order to avoid the 2f excitation force from exciting the resonance of the crankshaft rotor assembly, it is also necessary to ensure that F qz Satisfied: F qz >2(F max +10);

[0074] In summary, for a three-cylinder compressor, the natural frequency of the crankshaft rotor assembly should be set according to the maximum operating frequency. Specifically:

[0075] When F max ≤90Hz, then F qz Should meet: F qz >3(F max +10);

[0076] When F max >90Hz, then F qz Should meet: 2(F max +10)<F qz <3(Fmax -10).

[0077] Specifically, if Figure 5 、 Figure 6 、 Figure 7 As shown, the three types of compressors involved in the present invention generate gas torques that vary with the crankshaft angle during operation; the single-cylinder compressor has only one cylinder, and the gas torque fluctuates from 0-100%. The two-cylinder compressor has a phase difference of 180° between the gas torques of the two cylinders, and its gas torque fluctuates from 40% to 100%, and its gas pulsation is lower than that of the single-cylinder compressor. The three-cylinder compressor has a phase difference of 120° between the gas torques of the three cylinders, and its gas torque fluctuates from 60% to 100%, and the pulsation is further reduced.

[0078] Furthermore, if Figure 8 、 Figure 9 、 Figure 10 As shown in FIG, the magnitude relationship of the 1f, 2f, and 3f excitation forces generated when the three compressors are in operation is different. The gas torques of different compressor structures are different, resulting in different magnitude relationships of the 1f, 2f, and 3f excitation forces.

[0079] Furthermore, Figure 11 The figure shows that for a single-cylinder compressor, when the natural frequency of the crankshaft rotor assembly is relatively low, the crankshaft rotor assembly resonates with the excitation force of 1f within the maximum operating frequency range of the compressor and generates a relatively large acceleration.

[0080] Furthermore, Figure 12 It is shown that if the natural frequency of the crankshaft rotor assembly is appropriately increased, the 1f excitation force can be avoided from exciting the resonance of the crankshaft rotor assembly within the maximum operating frequency range of the compressor.

[0081] Furthermore, Figure 13 、 Figure 14 The following are two schemes and their effects for increasing the natural frequency of the crankshaft rotor assembly in a single-cylinder compressor; Figure 13 The first solution is shown: increasing the natural frequency of the crankshaft rotor to a point where the crankshaft rotor assembly cannot be resonated by the 1f excitation force, but can be resonated by the 2f excitation force. Since the 2f excitation force is relatively small in a single-cylinder compressor, although it is resonated by the 2f excitation force, the acceleration at resonance is small due to the small 2f excitation force (the acceleration generated by the 2f excitation force to excite the resonance of the crankshaft rotor assembly is 40% of the acceleration generated by the 1f excitation force to excite the resonance of the crankshaft rotor assembly). It is only necessary to shield some frequency points to obtain a better vibration and noise reduction effect. Figure 14 The second solution is shown: increasing the natural frequency of the crankshaft rotor assembly to a point where the crankshaft rotor assembly cannot be excited to resonance by the 2f excitation force.

[0082] Furthermore, Figure 15The following shows the natural frequency setting scheme for the crankshaft rotor assembly when the displacement of a two-cylinder compressor is less than 36cc: Since the displacement is small at this time, the 2f excitation force is also correspondingly small. Therefore, the natural frequency of the crankshaft rotor assembly can be appropriately reduced to a range within the maximum operating frequency where the 2f excitation force can excite the crankshaft rotor assembly to resonate. Due to the small 2f excitation force, the corresponding acceleration caused by the resonance of the crankshaft rotor assembly is also smaller than that of a two-cylinder compressor with a larger displacement (>36cc).

[0083] Furthermore, Figure 16 The following shows the setting scheme of the natural frequency of the crankshaft rotor assembly when the displacement of a two-cylinder compressor is greater than 36cc: the natural frequency is set so that the crankshaft rotor assembly cannot be excited to resonate by the 2f excitation force within the maximum operating frequency range, but can be excited to resonate by the 2f excitation force because the 3f excitation force is smaller than the 2f excitation force ( Figure 9 As shown in Figure 2, the acceleration generated by the resonance of the crankshaft rotor assembly caused by the 3f excitation force is also relatively small.

[0084] Furthermore, Figure 17 The setting scheme of the natural frequency of the crankshaft rotor assembly of the three-cylinder compressor is shown when the maximum operating frequency is less than 90Hz: within the maximum operating frequency range, the crankshaft rotor assembly cannot be excited to resonate by the 3f excitation force; further, Figure 18 The setup presented here is such that the crankshaft-rotor assembly can be excited to resonance by a 3f excitation force within the maximum operating frequency range.

[0085] Combining the above solutions, such as Figures 1 to 3 As shown, the natural frequency F of the crankshaft rotor assembly 3 qz Implementation of design values: Single-cylinder compressors, two-cylinder compressors, and three-cylinder compressors have differences in cost, structure, vibration characteristics, displacement, etc., and the corresponding crankshaft diameters and rotor masses are also significantly different, which are manifested in the following ways:

[0086] (1) Single-cylinder compressors have the lowest cost and are smaller in size, but the gas torque fluctuation is greater. They are generally used for models with 2 HP or less (compressor displacement is less than 16cc). The displacement is small, and the corresponding crankshaft diameter D and rotor assembly mass m are small. Therefore, in order to meet the maximum operating frequency of 120Hz, the natural frequency F of the crankshaft rotor assembly is qz Requirements (140Hz <F qz <220Hz or F qz >260Hz), according to the formula The suitable range of the crankshaft diameter D and rotor assembly mass m of a single-cylinder compressor should be: 10mm≤D≤14mm, 0.4kg≤m≤0.85, or 14 <D,m≤0.4kg;

[0087] (2) The gas torque fluctuation of the twin-cylinder compressor is smaller than that of the single-cylinder compressor, and the vibration is smaller. It can not only fully cover the capacity range of the single-cylinder compressor but also output a greater capacity. However, due to cost and vibration considerations, twin-cylinder compressors are usually mainly used in models with more than 2 HP (displacement greater than 20cc). For displacements between 20cc and 36cc, in order to meet the natural frequency F of the crankshaft rotor assembly, qz <2(F max -10), according to the formula The suitable range of crankshaft diameter D and rotor mass m for a two-cylinder compressor within this displacement range should be: D ≤ 16 mm, m > 0.7 kg; for models with a displacement greater than 36 cc, in order to make the natural frequency of the crankshaft rotor assembly satisfy formula 2 (F max +10)<F qz <3(F max -10), the appropriate range of crankshaft diameter D and rotor assembly mass m for a two-cylinder compressor within this displacement range should be: 17mm <D≤24mm,0.7kg<m<2.5kg;

[0088] (3) There are two main applications for three-cylinder compressors. One is to further expand the displacement range of two-cylinder compressors, and the other is to apply to special structural models (such as two-stage compressors). Therefore, it is mainly used in occasions where the displacement demand is large. According to the difference in its maximum operating frequency, its crankshaft diameter D and rotor assembly mass m should also be different. Specifically: for the compressor with the maximum operating frequency meeting F max ≤90Hz and F qz >3(F max +10) requirements, according to the formula The appropriate range of crankshaft diameter D and rotor assembly mass m should be: 23mm≤D, m≥1.6kg; and the maximum operating frequency of the compressor should meet 90<F max And 2(F max +10)<F qz <3(F max -10) requirements, according to the formula The appropriate range of crankshaft diameter D and rotor assembly mass m should be: 17mm≤D<23mm, 0.8kg <m<2.6kg。

[0089] Combined with the above scheme, when the compressor is still in the development and design stage, the above method can be used to determine and adjust the natural frequency F of the crankshaft rotor assembly in advance. qz, crankshaft diameter D, and rotor assembly mass m to achieve a better noise reduction effect; however, when the compressor has been finalized and mass-produced, if it is found during use that the resonance of the crankshaft-rotor assembly affects the low-frequency noise of the compressor, it will be very difficult to adjust the crankshaft diameter D or the rotor assembly mass m. The reason is that the above adjustments may require redesigning the tooling and measuring instruments for producing the compressor, which not only requires a large investment but also a long preparation time. For this reason, for compressors that are already in the production stage, an annular groove can be set at a specific position on the crankshaft. The diameter d of the annular groove can be adjusted according to Design verification is performed; the annular groove is located in the area between the rotor assembly and the upper flange (such as Figure 4 As shown), and the corresponding dimension value is set according to the compressor structure. For example, if the compressor is a single-cylinder compressor, the annular groove diameter value d (i.e., the crankshaft diameter in the annular groove) can be set according to the natural frequency value range of the crankshaft rotor assembly in the single-cylinder compressor. If the compressor is a two-cylinder compressor, the annular groove diameter value d (i.e., the crankshaft diameter in the annular groove) can be set according to the natural frequency value range of the crankshaft rotor assembly in the two-cylinder compressor. If the compressor is a three-cylinder compressor, the annular groove diameter value d (i.e., the crankshaft diameter in the annular groove) can be set according to the natural frequency value range of the crankshaft rotor assembly in the three-cylinder compressor.

[0090] The present invention proposes a compressor, which verifies and determines the factors affecting the natural frequency of the crankshaft rotor assembly through experiments, sets the optimal frequency range of the natural frequency of the crankshaft rotor assembly, and proposes specific parameter settings of the crankshaft to reduce the low-frequency vibration of the compressor, thereby solving the low-frequency noise of the existing compressor and reducing the low-frequency noise generated during the operation of the compressor.

[0091] It is further understood that in the present disclosure, "plurality" refers to two or more than two, and other quantifiers are similar. "And / or" describes the association relationship of associated objects, indicating that three relationships may exist. For example, A and / or B can mean: A exists alone, A and B exist at the same time, and B exists alone. The character " / " generally indicates that the related objects before and after are in an "or" relationship. The singular forms "a", "the" and "the" are also intended to include the plural forms, unless the context clearly indicates otherwise.

[0092] It will be further understood that the terms "first," "second," and the like are used to describe various types of information, but such information should not be limited to these terms. These terms are used solely to distinguish information of the same type from one another and do not indicate a particular order or level of importance. In fact, the terms "first," "second," and the like are fully interchangeable. For example, first information could be referred to as second information, and similarly, second information could be referred to as first information without departing from the scope of this disclosure.

[0093] It is further understood that although operations are described in a particular order in the drawings in the embodiments of the present disclosure, this should not be construed as requiring that the operations be performed in the particular order shown or in a serial order, or that all of the operations shown be performed to obtain the desired results. In certain circumstances, multitasking and parallel processing may be advantageous.

[0094] Other embodiments of the present disclosure will readily occur to those skilled in the art after considering the specification and practicing the invention disclosed herein. This application is intended to cover any variations, uses, or adaptations of the present disclosure that follow the general principles of the present disclosure and include common knowledge or customary techniques in the art not disclosed herein. The description and examples are to be considered as exemplary only, with the true scope and spirit of the present disclosure being indicated by the following claims.

[0095] It should be understood that the present disclosure is not limited to the exact structures described above and shown in the drawings, and that various modifications and changes may be made without departing from the scope thereof. The scope of the present disclosure is limited only by the scope of the appended claims.

Claims

1. A compressor, characterized in that: The compressor comprises a crankshaft rotor assembly (3), wherein the crankshaft rotor assembly (3) comprises a rotor (6) and a crankshaft (7); The natural frequency F of the crankshaft rotor assembly (3) qz Designed according to the following formula: The λ is a conventional coefficient, the D is the shaft diameter of the crankshaft (7), and the m is the mass of the crankshaft rotor assembly (3); The value range of λ is 0.75-0.

85.

2. The compressor according to claim 1, characterized in that When the compressor is a single-cylinder compressor, the natural frequency F qz satisfy: 20+F max <F qz <2(F max -10) or F qz >2(F max +10), the F max The maximum operating frequency of the compressor.

3. The compressor according to claim 2, characterized in that When the maximum operating frequency of the compressor F max At 120Hz, the F qz Meets: 140Hz <F qz <220Hz or F qz >260Hz.

4. The compressor according to claim 1, characterized in that The compressor is a two-cylinder compressor; when the total displacement of the compressor is less than 36cc, the natural frequency F qz Satisfied: F qz <2(F max -10); When the total displacement of the compressor is greater than 36cc, the natural frequency F qz Satisfaction: 2(F max +10) <F qz <3(F max -10).

5. The compressor according to claim 1, characterized in that The compressor is a three-cylinder compressor; when the maximum operating frequency of the compressor is F max ≤90Hz, the natural frequency F qz Satisfied: F qz >3(F max +10); when the maximum operating frequency of the compressor F max >90Hz, the natural frequency F qz Satisfaction: 2(F max +10) <F qz <3(F max -10).

6. The compressor according to claim 1, characterized in that The value ranges of the shaft diameter D of the crankshaft (7) and the mass m of the crankshaft rotor assembly (3) are: 10mm≤D≤14mm, 0.4kg≤m≤0.85, or 14 <D,m≤0.4kg。 7. The compressor according to claim 4, characterized in that When the total displacement of the compressor is less than 36cc, the range of the shaft diameter D of the crankshaft (7) and the mass m of the crankshaft rotor assembly (3) is: D≤16mm, m>0.7kg; when the total displacement of the compressor is greater than 36cc, the range of the shaft diameter D of the crankshaft (7) and the mass m of the crankshaft rotor assembly (3) is: 17mm <D≤24mm,0.7kg<m<2.5kg。 8. The compressor according to claim 5, characterized in that When the maximum operating frequency of the compressor F max ≤90Hz, the range of the diameter D of the crankshaft (7) and the mass m of the crankshaft rotor assembly (3) is: 23mm≤D, m≥1.6kg; when the maximum operating frequency F of the compressor is max When the frequency is >90 Hz, the range of the diameter D of the crankshaft (7) and the mass m of the crankshaft rotor assembly (3) is: 17 mm ≤ D < 23 mm, 0.8 kg <m<2.6kg。 9. The compressor according to claim 1, characterized in that The crankshaft (7) is provided with an annular groove (71), and the annular groove (71) is located between the rotor assembly (3) and the upper end surface of the upper flange; the compressor adjusts the natural frequency F of the crankshaft rotor assembly (3) by the diameter value d of the annular groove (71). qz The natural frequency F qz satisfy:

10. The compressor according to any one of claims 1 to 9, characterized in that The compressor is a rolling rotor type compressor.

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