The structure and design method of the suction port of a twin-screw compressor

Abstract

A structure and design method for a twin-screw compressor suction port, comprising an axial suction port and a radial suction port, wherein a distance is provided between the end face of the axial suction port and the end face of the suction-side bearing; the axial suction port is provided with a delay angle, so that the suction closing angle is delayed; the radial suction port is shortened accordingly according to the delay angle of the axial suction port, so that the axial distance of the radial suction port does not exceed the theoretical suction spiral corresponding to the suction delay. The twin-screw compressor suction port structure designed by the present invention fully considers the uneven pressure distribution characteristics inside the rotor cavity, avoids the problem of pressure mismatch at local ports, and can achieve sufficient air intake, thereby effectively improving the volumetric efficiency and adiabatic efficiency of the compressor; in addition, the present invention can be used to improve the performance of existing compressor products. Without changing the rotor, only the suction port on the casing needs to be modified, which is convenient to process, low in cost, and easy to implement.

Description

Technical Field

[0001] The invention belongs to the field of refrigeration compressors, and in particular relates to a structure of an air suction port of a twin-screw compressor and a design method thereof. Background Art

[0002] Improving the efficiency of industrial electricity consumption is imperative for environmental protection and energy conservation and emission reduction. Compressors, as core components of various power machinery, account for approximately 15% of industrial electricity consumption. Twin-screw refrigeration compressors, as rotary compressors, are commonly used in medium-sized refrigeration and heat pump units due to their long life, high reliability, and high efficiency. Even a 1% improvement in twin-screw compressor performance can save significant energy.

[0003] Current improvements to compressors primarily focus on optimizing rotor profiles and improving seals. For refrigeration compressors, liquid injection can significantly improve sealing, reduce leakage, and cool the compressed gas, leaving little room for improvement. Orifice design, on the other hand, primarily focuses on suppressing and eliminating exhaust pulsation, with relatively few improvements to intake orifices. Currently, intake orifice design varies widely among manufacturers, lacking clear design guidelines.

[0004] Furthermore, current compressor design and optimization are based on a working chamber model, assuming the same pressure within each screw tooth. However, pressure within the compressor propagates at the speed of sound, and for refrigerants with lower sound velocities, this pressure propagation rate is limited. For common refrigerants such as R134a, the sound velocity is typically 150 m / s, which further decreases to approximately 130 m / s after oil injection. When the compressor rotates 10° at 3000 rpm, the sound wave only travels approximately 72 mm, far less than the length of the spiral. Therefore, for larger compressors, the pressure within the rotor during compression is far from balanced. When pressure distribution is ignored during orifice design, the uneven pressure distribution within the rotor cavity can lead to pressure mismatches at local orifices. This mismatch can lead to reduced efficiency, and therefore, the current orifice design process lacks consideration of the compressor's internal characteristics. Summary of the Invention

[0005] The purpose of the present invention is to address the problem that the design of the suction orifice in the above-mentioned prior art ignores the uneven pressure distribution inside the rotor cavity, and to provide a structure and design method for the suction orifice of a twin-screw compressor, which can significantly improve the volume flow rate and insulation efficiency of the refrigeration compressor without changing the rotor.

[0006] In order to achieve the above object, the present invention has the following technical solutions:

[0007] A structure of a twin-screw compressor suction orifice includes an axial suction orifice and a radial suction orifice, wherein a distance is set between the end face of the axial suction orifice and the end face of the suction-side bearing; the axial suction orifice is set with a delay angle so that the suction closing angle is delayed; the radial suction orifice is shortened accordingly according to the delay angle of the axial suction orifice, so that the axial distance of the radial suction orifice does not exceed the theoretical suction spiral line corresponding to the suction delay.

[0008] As a preferred solution, the distance between the end face of the axial suction port and the end face of the suction side bearing is set according to the following calculation expression:

[0009]

[0010] Where, d s v is the distance between the suction side bearing end face and the suction end face, in mm; s is the average inspiratory flow velocity at the inspiratory orifice cross section, in m / s; k s is the resistance characteristic coefficient; ρ s is the refrigerant suction density, in kg / m 3 .

[0011] Furthermore, the resistance characteristic coefficient k s Take 0.006.

[0012] As a preferred solution, the delay angle of the axial air intake orifice is set according to the following calculation expression:

[0013]

[0014] Where Δα s is the axial suction orifice delay angle, in degrees; n is the rotation speed, in rpm; L s is the length of the spiral, in m; k f is the reentry coefficient; c is the refrigerant sound velocity, in m / s.

[0015] Furthermore, for twin-screw compressors with gear ratios of 5 / 6 and 6 / 7, the foldback coefficient k f Take 4.

[0016] As a preferred solution, when the radial intake orifice is shortened accordingly according to the delay angle of the axial intake orifice, when the average intake flow rate does not exceed 30 m / s, the radial intake orifice is completely closed or the minimum radial orifice depth is selected; when the average intake flow rate exceeds 30 m / s, the radial intake orifice is opened and the maximum axial distance of the radial intake orifice is limited. The maximum axial distance of the radial intake orifice is determined by the actual delay angle of the axial intake orifice.

[0017] Furthermore, the calculation expression for determining the maximum axial distance of the radial air intake orifice by the actual delay angle of the axial air intake orifice is:

[0018]

[0019] Where k L is the radial orifice length coefficient; L0 is the theoretical radial orifice length, in m; d0 is the redundant length; k d is the delay coefficient; c is the refrigerant sound velocity, in m / s; n is the speed, in rpm; Δα s is the axial suction orifice delay angle, in degrees.

[0020] Furthermore, the redundant length d0 is determined according to the radial inspiratory flow rate, and the radial orifice opening length is greater than or equal to the redundant length d0; the delay coefficient k d Adjusted according to the number of rotor teeth.

[0021] A method for designing the structure of a twin-screw compressor suction port, comprising:

[0022] The air flow sucked into the compressor is blocked by the axial suction bearing. By increasing the distance between the axial suction bearing and the end face of the axial suction orifice, the dynamic pressure loss of the axial velocity is reduced, so that the flow direction of the air flow after entering the rotor working chamber is still mainly axial.

[0023] By delaying the suction of the axial suction orifice, the actual suction capacity of the screw compressor is made higher than the theoretical suction capacity;

[0024] The axial length of the radial air intake opening is shortened accordingly according to the delay angle of the axial air intake opening.

[0025] As a preferred solution, in the step of delaying the suction of the axial suction orifice to make the actual suction volume of the screw compressor higher than the theoretical suction volume, after reaching the theoretical suction contact position, suction is continued through the delayed axial orifice; after the suction is completely closed, the dynamic pressure in the rotor cavity is converted into static pressure, the average pressure is increased, the work done by the twin-screw compressor is reduced, and the insulation efficiency is improved.

[0026] Compared with the prior art, the present invention has at least the following beneficial effects:

[0027] The twin-screw compressor suction orifice structure designed by the present invention fully considers the uneven pressure distribution characteristics inside the rotor cavity. By setting a distance between the axial suction orifice end face and the suction side bearing end face, a delay angle is set for the axial suction orifice, so that the suction closing angle is delayed, and the radial suction orifice is shortened accordingly according to the delay angle of the axial suction orifice, thereby avoiding the problem of pressure mismatch at the local orifice, and achieving sufficient air intake, thereby effectively improving the volumetric efficiency and insulation efficiency of the compressor; in addition, the suction orifice structure and design method provided by the present invention can be used to improve the performance of existing compressor products. Without changing the rotor, only the suction orifice on the casing needs to be modified. It is easy to process, low in cost, and easy to implement. BRIEF DESCRIPTION OF THE DRAWINGS

[0028] Figure 1 Schematic diagram of the suction hole cavity of the suction hole structure of the twin-screw compressor according to an embodiment of the present invention;

[0029] Figure 2 Schematic diagram of the actual casing structure of the suction port structure of the twin-screw compressor according to the embodiment of the present invention;

[0030] Figure 3 A schematic diagram of a planar expansion of the suction orifice structure of a twin-screw compressor according to an embodiment of the present invention;

[0031] Figure 4 Simulation results and principle diagram of the structural design method of the twin-screw compressor suction port according to the embodiment of the present invention;

[0032] Figure 5 Velocity vector diagram of the radial orifice of the suction orifice structure of the twin-screw compressor according to the embodiment of the present invention;

[0033] Figure 6 A diagram of the turbulence dissipation rate in the radial orifice of the suction orifice structure of the twin-screw compressor according to an embodiment of the present invention;

[0034] In the figure: 1- radial suction surface; 2- axial suction surface; 3- radial suction closed theoretical spiral; 4- axial suction closed theoretical orifice; 5- axial suction closed actual orifice; 6- axial suction delay angle Δα s ;7-suction bearing end face;8-suction bearing seat and suction end face distance d s ;9-maximum axial position of radial suction;10-maximum axial distance d of radial suction max ;11-depth of radial suction cavity;12-radial suction orifice;13-suction side bearing;14-axial suction baffle;15-intersection line of yin and yang rotor holes;21-high pressure area of ​​exhaust end face of working cavity;22-low pressure area of ​​suction end face of working cavity;23-velocity vector of axial suction delay area;24-gas surge in radial suction cavity;25-gas turbulence dissipation rate in radial suction cavity. DETAILED DESCRIPTION

[0035] The present invention will be further described in detail below with reference to the accompanying drawings and embodiments.

[0036] See also Figure 1 The embodiment of the present invention proposes a structure for the suction port of a twin-screw compressor. Compared with the traditional compressor port design, the axial suction port end face maintains a certain distance from the suction side bearing end face, which increases the depth of the axial suction cavity, increases the axial intake surface 2, and reduces the radial intake surface 1. The suction structure should have a complete axial suction port, and the axial suction surface 2 should be sufficiently far away from the suction bearing end face 7 so that the suction flow can maintain an axial state and reduce the suction resistance. The distance d between the suction bearing seat and the suction end face is s By controlling Figure 2 The distance is ensured by the axial length of the axial air intake baffle 14 shown in FIG. Figure 3 As shown, the distance d between the suction bearing seat and the suction end face is s Calculated by the following formula:

[0037]

[0038] Where, d s v is the distance between the suction side bearing end face and the suction end face, in mm; s is the average inspiratory flow velocity at the inspiratory orifice cross section, in m / s; k s is the resistance characteristic coefficient, preferably 0.006; ρ s is the refrigerant suction density, in kg / m 3 .

[0039] The axial suction surface 2 has a certain delay angle, and the position of the axial orifice is delayed from the axial suction closed theoretical orifice 4 to the axial suction closed actual orifice 5. Axial suction delay angle Δα s A preliminary estimate is made using the following formula:

[0040]

[0041] Where Δα s is the axial inhalation delay angle, in degrees; n is the rotation speed, in rpm; L s is the length of the spiral, in m; k f is the foldback coefficient, preferably 4 for twin-screw compressors with 5 / 6 and 6 / 7 gear ratios; c is the refrigerant sound velocity, expressed in m / s. In a preferred embodiment, for a male rotor diameter of 212 mm, refrigerant R134a, a gear ratio of 5 / 6, and a speed of 3000 rpm, the maximum retardation angle can be selected to be 20°.

[0042] See also Figure 3 The radial orifice is reduced accordingly according to the delay angle of the axial air inlet orifice, and its axial distance should not exceed the theoretical air inhalation spiral line corresponding to the air inhalation delay. When the air inhalation orifice design method stipulates that the average air inhalation flow rate does not exceed 30m / s, the radial air inhalation orifice can be completely closed; when the average air inhalation flow rate exceeds 30m / s, the radial air inhalation orifice is opened, but its maximum axial distance needs to be limited. The maximum axial distance d of radial air inhalation max Determined by the actual delay angle:

[0043]

[0044] Where k L is the radial orifice length coefficient, preferably 0.8 when the twist angle is 300 degrees; L0 is the theoretical radial orifice length, in meters; d0 is the redundant length, which needs to be determined according to the radial inhalation flow rate, preferably 0.02m. The final radial orifice opening length should be greater than or equal to d0; k d is the delay coefficient, preferably 1.2, which should be adjusted according to the number of rotor teeth.

[0045] Another embodiment of the present invention further provides a method for designing the structure of a twin-screw compressor suction port, comprising:

[0046] The airflow sucked into the compressor first passes through the obstruction of the axial suction side bearing 13. By increasing the distance between the suction side bearing 13 and the axial suction surface 2, the dynamic pressure loss of the axial velocity can be reduced. After the airflow enters the rotor working chamber, its flow direction is still mainly axial. As shown in the high-pressure area 21 of the exhaust end face of the working chamber, the pressure distribution of the rotor chamber during the suction process is uneven, and an obvious pressure gradient is shown in the cloud diagram. Before the suction is closed, the high-pressure area 21 of the exhaust end face of the working chamber has risen, but due to the limited speed of pressure propagation, the pressure of the low-pressure area 22 of the suction end face of the working chamber is still relatively low. Therefore, after reaching the theoretical suction contact position, it can still continue to be sucked in through the axial suction delay area velocity vector 23. By delaying the suction, the actual suction volume of the screw compressor can be made higher than the theoretical suction volume, thereby improving the volumetric efficiency. And after the suction is completely closed, the dynamic pressure in the rotor chamber is converted into static pressure, the average pressure is increased, the work done by the compressor is reduced, and the adiabatic efficiency is improved.

[0047] The high-pressure area 21 at the exhaust end face of the working chamber may leak through the radial suction orifice, so the radial orifice length coefficient is set to shorten it. In addition, as shown by the gas surge 24 in the radial suction chamber, radial suction will cause the gas in the working chamber to be thrown out by the rotor and re-inhaled, and continue to surge. On the one hand, it aggravates the internal friction of the gas. As shown by the gas turbulence dissipation rate 25 in the radial suction chamber, the gas dissipates as internal energy in the radial orifice, so the depth of the radial orifice should be reduced to suppress the gas surge. On the other hand, the radial suction orifice will make the pressure in the rotor cavity more balanced, disrupting the overall flow of the inhaled rotor gas along the spiral direction, resulting in a reduction in the axial delayed suction effect, so the axial length of the radial suction orifice is reduced accordingly to match the axial suction delay.

[0048] In view of the fact that the existing suction orifice design ignores the problem of uneven pressure distribution inside the rotor cavity, the embodiment of the present invention fully considers the uneven pressure distribution characteristics inside the rotor cavity when designing the suction orifice of the twin-screw compressor, avoiding the problem of pressure mismatch at the local orifice, and achieving sufficient air intake, thereby effectively improving the volumetric efficiency and insulation efficiency of the compressor; in addition, the suction orifice structure and design method provided by the embodiment of the present invention can be used to improve the performance of existing compressor products. Without changing the rotor, only the suction orifice on the casing needs to be modified. It is convenient to process, low in cost, and easy to implement.

[0049] The above-described embodiments are only used to illustrate the technical solutions of the present application, rather than to limit them. Although the present application has been described in detail with reference to the aforementioned embodiments, those skilled in the art should understand that they can still modify the technical solutions described in the aforementioned embodiments, or make equivalent replacements for some of the technical features therein. These modifications or replacements do not deviate the essence of the corresponding technical solutions from the spirit and scope of the technical solutions of the various embodiments of the present application, and should all be included in the scope of protection of the present application.

Claims

1. A structure of a twin-screw compressor suction port, characterized in that: It includes an axial suction orifice and a radial suction orifice, wherein a distance is set between the end face of the axial suction orifice and the end face of the suction-side bearing; the axial suction orifice is set with a delay angle so that the suction closing angle is delayed; the radial suction orifice is shortened accordingly according to the delay angle of the axial suction orifice, so that the axial distance of the radial suction orifice does not exceed the theoretical suction spiral line corresponding to the suction delay; The distance between the end face of the axial air intake port and the end face of the air intake side bearing is set according to the following calculation expression: Where, d s v is the distance between the suction side bearing end face and the suction end face, in mm; s is the average inspiratory flow velocity at the inspiratory orifice cross section, in m / s; k s is the resistance characteristic coefficient; ρ s is the refrigerant suction density, in kg / m 3 ; The delay angle of the axial air intake orifice is set according to the following calculation expression: Where Δα s is the axial suction orifice delay angle, in degrees; n is the rotation speed, in rpm; L s is the length of the spiral, in m; k f is the reentry coefficient; c is the refrigerant sound velocity, in m / s; When the radial suction orifice is shortened accordingly according to the delay angle of the axial suction orifice, when the average suction flow rate does not exceed 30 m / s, the radial suction orifice is completely closed or the minimum radial orifice depth is selected; when the average suction flow rate exceeds 30 m / s, the radial suction orifice is opened and the maximum axial distance of the radial suction orifice is limited. The maximum axial distance of the radial suction orifice is determined by the actual delay angle of the axial suction orifice. The calculation expression for determining the maximum axial distance of the radial air intake orifice by the actual delay angle of the axial air intake orifice is: Where k L is the radial orifice length coefficient; L0 is the theoretical radial orifice length, in m; d0 is the redundant length; k d is the delay coefficient; c is the refrigerant sound velocity, in m / s; n is the speed, in rpm; Δα s is the axial suction orifice delay angle, in degrees.

2. The structure of the suction port of the twin-screw compressor according to claim 1, characterized in that: The characteristic resistance coefficient k s Take 0.

006.

3. The structure of the suction port of the twin-screw compressor according to claim 1, characterized in that: For twin-screw compressors with gear ratios of 5 / 6 and 6 / 7, the foldback coefficient k f Take 4.

4. The structure of the suction port of a twin-screw compressor according to claim 1, characterized in that: The redundant length d0 is determined according to the radial inspiratory flow rate, and the radial orifice opening length is greater than or equal to the redundant length d0; the delay coefficient k d Adjusted according to the number of rotor teeth.

5. A method for designing the structure of the suction port of a twin-screw compressor according to any one of claims 1 to 4, characterized in that: include: The air flow sucked into the compressor is blocked by the axial suction bearing. By increasing the distance between the axial suction bearing and the end face of the axial suction orifice, the dynamic pressure loss of the axial velocity is reduced, so that the flow direction of the air flow after entering the rotor working chamber is still mainly axial. By delaying the suction of the axial suction orifice, the actual suction capacity of the screw compressor is made higher than the theoretical suction capacity; The axial length of the radial air intake opening is shortened accordingly according to the delay angle of the axial air intake opening.

6. The design method according to claim 5, characterized in that: In the step of delaying the suction of the axial suction orifice so that the actual suction volume of the screw compressor is higher than the theoretical suction volume, after reaching the theoretical suction contact position, suction is continued through the delayed axial orifice; after the suction is completely closed, the dynamic pressure in the rotor cavity is converted into static pressure, the average pressure is increased, the work done by the twin-screw compressor is reduced, and the insulation efficiency is improved.

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