Anti-rotation flow guide integrated radial seal based on dynamic and static pressure effects

Abstract

The anti-rotation flow guide integrated radial seal comprises an anti-rotation flow guide structure and a plurality of sealing teeth, the anti-rotation flow guide structure and the sealing teeth are sequentially arranged on the inner surface of a sealing stator in the flowing direction, the anti-rotation flow guide structure is installed at an inlet of the sealing stator, high-speed circumferential rotational flow caused by pre-rotation of an air inlet can be weakened, and the anti-rotation flow guide effect is improved. And the pressure distribution between the air inlet and the adjacent sealing teeth of the anti-rotation flow guide structure can be changed, so that the excitation is reduced, and the leakage rate is reduced. Wherein the spiral line type anti-rotation flow guide structure is composed of a plurality of spiral flow guide plates distributed in the circumferential direction, a millimeter-level spiral deep groove area is designed at an inlet, the flow pattern of airflow is changed through the hydrostatic pressure effect, the circumferential rotational flow velocity of the surface of the rotor is reduced, and the kinetic energy dissipation of the airflow is increased. Or the spiral line type anti-rotation flow guide structure is a boss with a spiral flow guide groove, a reasonable groove type structure is designed according to the rotation direction of the rotor and the working condition so as to change the pressure distribution of the airflow, and the throttling and flow guide effects are achieved.

Description

Technical Field

[0001] The present invention belongs to the technical field of rotating machinery, and particularly relates to an integrated anti-rotation and flow-guiding radial seal based on hydrostatic and hydrodynamic effects, which is a radial seal with an integrated anti-rotation and flow-guiding structure provided at the fluid inlet. Background Art

[0002] The sealing device is one of the important components on the shaft of rotating machinery such as steam turbines, compressors, and various pumps. It is usually installed at the gap between rotating components such as rotors and blade tips and stationary components such as cylinders and diaphragms. When the fluid flows from the high-pressure end to the low-pressure end, the sealing device plays a role in preventing fluid leakage, greatly improving the efficiency of the material or energy conversion of the rotating machinery. Researchers have carried out a large number of structural innovations on axial sealing devices such as mechanical seals and dry gas seals, and the sealing performance has been greatly improved by using the hydrodynamic effect and pumping effect. For example, the invention patent with the patent application number CN 202311005963.6, whose name is "A double-step contraction type spiral groove gas film sealing structure". This structure divides the spiral groove into three parts. The first spiral groove part is divided into three-level step contraction structures, and the second spiral groove part is divided into two-level step contraction structures, effectively enhancing the blocking effect and pressure boosting effect of the airflow and reducing the seal leakage. At present, there is less research on the application of the hydrodynamic effect and pumping effect in radial sealing devices, so new radial seal structures need to be developed.

[0003] Typical radial sealing devices include labyrinth seals, honeycomb seals, hole-type seals, etc. The labyrinth seal is widely used due to its advantages such as simple structure, convenient installation, low cost, and easy maintenance. However, during operation, vortices are extremely likely to be generated in the seal cavity, which in turn causes unsteady airflow excitation force, posing a certain threat to the safe and stable operation of the rotor. Changing such a sealing structure to achieve the limitation of vortices and the reduction of the pressure change at the inlet and outlet of the seal is of great innovative value for improving its engineering performance, and many researchers have carried out relevant research.

[0004] The utility model patent with the application number CN 201621276680.0 and the name "A reverse swirl honeycomb seal structure with spiral grooves" discloses a new type of honeycomb seal structure. It opens spiral grooves in the traditional honeycomb belt in the opposite direction of the rotor rotation direction, reducing the circumferential airflow formed by the rotor rotation and increasing the rotor stability. The design of the height difference of the spiral grooves in this structure lacks theoretical guidance and does not consider the influence of the spiral grooves on the pressure distribution in the seal cavity. Another example is that Chen Huilong et al. disclosed an annular seal assembly that can effectively improve the rotor dynamic characteristics (application number: CN 201721532030.2). Its main structure is to open multiple rows of spiral grooves and circumferential labyrinth grooves on the outer surface of the rotor or the inner surface of the seal sleeve, using the dynamic pressure effect and pumping effect of the spiral grooves to improve the pressure distribution in the seal area, improve the rotor centering performance, and enhance the rotor operation stability. This structure does not consider the influence of the spiral grooves on the circumferential swirl in the seal cavity and does not describe in detail the action mechanism of the dynamic pressure effect and pumping effect.

[0005] In the research on the design of labyrinth seal structures, some new labyrinth seal structures have been disclosed in the existing patent materials.

[0006] Liu Yan et al. disclosed a reverse swirl steam seal device for preventing steam turbine excitation (application number: CN 200520041731.7) and gave a new structure for arranging flow guides between the seal teeth. Its main structure is to install pre-swirl flow guides and swirl grooves on the existing structure of the staggered labyrinth seal, so that the airflow forms a reverse swirl, reducing the influence of airflow excitation. This structure has certain requirements for the axial dimension of the steam seal, and the invention uses linear flow guides, which can only change the airflow pattern and cannot change the airflow pressure distribution. In addition, machining swirl grooves at the root of the seal teeth will not only increase the leakage amount but also may affect the strength of the seal teeth.

[0007] The invention patent with the application number CN 201510292681.8 and the title "An anti-rotation plate structure with anti-rotation and vibration suppression functions" discloses an anti-rotation structure that changes the inlet flow rate. It arranges guide vanes on the inner surface of the stator ring upstream of the sealed inlet, so that the inner surface of the stator ring, the guide vanes, and the outer surface of the rotor form a variable cross-section annular fluid channel, thereby changing the inlet fluid flow rate, reducing the pre-rotation at the sealed inlet, and improving the rotor stability. In this invention, the influence of the change in the cross-sectional area of the fluid passage on the fluid pressure is not considered. Another example is the invention patent with the publication number CN 110080829 A and the title "An anti-rotation plate structure with holes and grooves on the inner surface", which discloses an anti-rotation plate with holes and grooves on the inner surface. This invention not only blocks and guides the circumferential air flow, but also has a dissipation effect similar to that of hole-type seals and honeycomb seals when the circumferential air flow passes through the anti-rotation plate, further suppressing the generation of circumferential swirl. However, in this invention, there is a lack of theoretical guidance for the design of the groove depth, and the situation where the groove depth forms static pressure or dynamic pressure effects and then affects the pressure distribution is not distinguished; moreover, the anti-rotation plates used are all linear anti-rotation plates, and the effective control of the inlet and outlet air flow pressures is poor.

[0008] The patent "A labyrinth seal for optimizing the dynamic characteristics of a rotor based on an anti-rotation plate between teeth" (CN 113606344 A) discloses a labyrinth seal with an anti-rotation plate installed between teeth to optimize the dynamic characteristics of the entire rotor system; this patent deflects the air flow passing through the anti-rotation plate in the reverse rotation direction, thereby reducing the circumferential swirl in the sealed cavity and further enhancing the seal stability. Compared with the circumferential swirl in the sealed cavity, the influence of the inlet air flow on the seal stability is more significant, but this seal structure does not consider it.

[0009] On the other hand, regarding the research on new structures of labyrinth seals, a structural modification plan for the seal teeth has also been proposed. For example, the invention patent with the publication number CN 108533757 A and the title "A labyrinth seal structure capable of increasing chamber dissipation" discloses eddy current small teeth arranged on the inner wall surface of the stator, both sides of the seal teeth, and the surface of the boss. The eddy current small teeth are arranged along the reverse air flow direction, playing a role in blocking the air flow and thus reducing the leakage; the eddy current small tooth structure can change the generation of eddy current flow in the sealed cavity and can dissipate the air flow eddy current; however, the structural parameters need to be designed, and this design method is relatively complex.

[0010] For another example, the invention patent with the application number CN 202110486037.X and the title "Self-stopping swirl labyrinth seal based on tip fin structure and its manufacturing method" discloses a tip fin arranged at the tip of the sealing tooth. Since its deflection direction is opposite to the rotation direction of the rotor, violent eddies will be generated when the air flow passes through the tip fin, increasing the kinetic energy dissipation of the air flow in the sealing cavity and reducing the leakage rate. The tip fin of this structure is located at the tip of the sealing tooth. To ensure the sealing effect, the gap between the sealing tooth and the rotor surface is small. When the air flow pressure is high, the tip fin may deform and rub against the rotor. In addition, while changing the swirl flow pattern, this structure fails to analyze the change in fluid pressure, which may lead to changes in local pressure, and the local pressure increase may excite the vibration of the rotor system.

[0011] In summary, for the above two types of diversion structures, they both start from the perspective of changing the fluid flow pattern to reduce the generation of circumferential swirl. However, the impact of the above structures on the rotor stability is not clear. In addition, the above existing technologies do not consider the regulating effect of the diversion structure on the fluid pressure and do not explain the mechanism required for the diversion structure to change the flow pattern. Therefore, it is necessary to develop a solution and a new structure with higher stability and high sealing performance based on the hydrodynamic and hydrostatic effects. Summary of the Invention

[0012] In order to overcome the above-mentioned disadvantages of the prior art, the purpose of the present invention is to provide a swirl-preventing and diversion-integrated radial seal based on hydrodynamic and hydrostatic effects to achieve the purpose of reducing excitation and leakage rate.

[0013] In order to achieve the above purpose, the technical solution adopted by the present invention is:

[0014] A swirl-preventing and diversion-integrated radial seal based on hydrodynamic and hydrostatic effects includes a swirl-preventing and diversion structure and several sealing teeth sequentially arranged on the inner surface of the seal stator along the flow direction, wherein the swirl-preventing and diversion structure is installed at the inlet of the seal stator.

[0015] In one embodiment, there is a radial radius gap C1 between the surface of the sealing tooth and the rotor, and the calculation formula is: C1 = (0.1 - 1)d / 1000. The calculation formula for H1 of the sealing tooth is H1 = [(d1 - d) / 2] - C1, where d1 is the diameter of the seal stator and d is the diameter of the rotor. The calculation formula for the axial gap L1 between adjacent sealing teeth is L1 = (8 - 36)C1.

[0016] In one embodiment, the height H of the swirl-preventing and diversion structure is H = H1, so that C = C1, where C is the radial radius gap between the swirl-preventing and diversion structure and the rotor surface; the axial gap between the swirl-preventing and diversion structure and the adjacent sealing tooth is L, and L = (1 / 5 - 1 / 2)L1; the axial dimension of the swirl-preventing and diversion structure is B, and B = (1 - 3)L1.

[0017] In one embodiment, the anti-rotation and flow-guiding structure is a spiral anti-rotation and flow-guiding structure.

[0018] In one embodiment, the polar coordinate form of the logarithmic spiral of the spiral anti-rotation and flow-guiding structure is: r = aexp(θcotβ), where r is the polar radius; a is the starting polar radius, that is, the axial dimension of the sealing stator; β is the spiral angle, and its value is the angle formed by the tangent at any point on the spiral and the line connecting the center of the circle to that point; θ is the circumferential angle between the logarithmic spiral at the polar radius and the starting point, and the parametric coordinate form is: where b is the exponential factor, and its value range is 0.25 - 0.4.

[0019] In one embodiment, the spiral anti-rotation and flow-guiding structure is composed of a plurality of spiral flow-guiding plates distributed circumferentially. The height of the spiral flow-guiding plate is the height H of the anti-rotation and flow-guiding structure. The axial dimension of the spiral flow-guiding plate and the axial clearance with the adjacent sealing teeth are both consistent with the anti-rotation and flow-guiding structure, and the structure adopts the logarithmic spiral of the spiral anti-rotation and flow-guiding structure.

[0020] In one embodiment, the height of the spiral flow-guiding plate is in millimeters, so as to form a spiral deep groove area in millimeters at the inlet of the sealing stator, meeting the static pressure effect condition; and the spiral direction of the spiral flow-guiding plate is opposite to the rotation direction of the rotor, so that the circumferential swirl flow on the rotor surface flows towards the inlet.

[0021] In one embodiment, the spiral anti-rotation and flow-guiding structure is a boss with spiral flow-guiding grooves, which is composed of a stator boss distributed circumferentially and a number of spiral flow-guiding grooves evenly distributed circumferentially on the inner surface of the stator boss. The height of the stator boss is the height H of the anti-rotation and flow-guiding structure.

[0022] In one embodiment, the height of the stator boss is in millimeters, and spiral flow-guiding grooves with a height in micrometers are opened on its inner surface, thereby forming a spiral shallow groove area in micrometers at the inlet to form a dynamic pressure effect. The groove depth h of the spiral flow-guiding groove is (0.001 - 0.01)C1 to meet the formation condition of the main dynamic pressure effect of the fluid, where C1 is the radial radius clearance between the sealing teeth and the surface of the rotor.

[0023] In one embodiment, when the rotation direction of the rotor is opposite to the spiral direction of the spiral flow-guiding groove, the air flow on the inlet side generates a pressure rise along the spiral flow-guiding groove, thereby forming a high-pressure sealing area at the end of the spiral flow-guiding groove to achieve a throttling effect; when the rotation direction of the rotor is the same as the spiral direction of the spiral flow-guiding groove, the air flow on the inlet side generates a pressure drop along the spiral flow-guiding groove, reducing the impact vibration caused by the high-pressure air flow to achieve a flow-guiding effect.

[0024] Compared with the prior art, the beneficial effects of the present invention are:

[0025] 1. By arranging an anti-rotation and flow-guiding structure at the inlet, the present invention can not only weaken the high-speed circumferential swirl caused by the pre-swirl at the air inlet, but also change the pressure distribution between the air inlet and the adjacent sealing teeth of the anti-rotation and flow-guiding structure, thereby reducing excitation and leakage.

[0026] 2. The spiral flow-guiding plate proposed by the present invention designs a millimeter-level spiral deep groove area at the inlet and meets the conditions of the hydrostatic effect. By the hydrostatic effect of the fluid, the flow pattern of the air flow is changed, the circumferential swirl velocity on the rotor surface is reduced, and the kinetic energy dissipation of the air flow is increased, thereby reducing the air flow excitation caused by the steam seal and improving the rotor stability.

[0027] 3. The boss with spiral flow-guiding grooves proposed by the present invention designs a micron-level spiral shallow groove area at the inlet and meets the formation conditions of the dynamic pressure effect, and works under the dynamic pressure effect of the fluid. According to the rotation direction of the rotor and the working conditions, a reasonable groove structure is designed to change the pressure distribution of the air flow, achieving the effects of throttling and flow guiding. Specifically, under high-pressure air flow conditions, the spiral direction of the spiral flow-guiding groove is designed to be the same as the rotation direction of the rotor to reduce the air flow pressure and the impact vibration caused by the air flow, achieving the flow guiding effect; under low-pressure air flow conditions, the seal can bear the impact of the air flow, so by designing the spiral direction of the flow groove to be opposite to the rotation direction of the rotor to increase the air flow pressure, forming a high-pressure seal area and reducing the leakage. BRIEF DESCRIPTION OF THE DRAWINGS

[0028] Figure 1 It is a two-dimensional sectional view of a labyrinth seal with a logarithmic spiral line type flow-guiding plate.

[0029] Figure 2 It is an axonometric view of a labyrinth seal with a spiral flow-guiding plate.

[0030] Figure 3 It is a developed view of a labyrinth seal with a spiral flow-guiding plate.

[0031] Figure 4 It is an axonometric view of a labyrinth seal with spiral flow-guiding grooves.

[0032] Figure 5 It is a partial enlarged view of the surface at a of a labyrinth seal with spiral flow-guiding grooves. DETAILED DESCRIPTION OF THE INVENTION

[0033] The following will describe the embodiments of the present invention in detail with reference to the drawings and embodiments.

[0034] The present invention is a radial seal based on the integrated anti-rotation and flow-guiding of the hydrostatic and dynamic pressure effects, specifically a labyrinth seal containing a special inlet anti-rotation and flow-guiding structure, to improve the pressure distribution of the inlet air flow, achieve the purpose of throttling or flow guiding, and improve the sealing performance.

[0035] Specifically, as Figure 1As shown in the figure, an integrated radial seal based on the hydrostatic and hydrodynamic effects for anti-rotation and flow guiding includes a seal stator 1, seal teeth 1-1, and an anti-rotation and flow guiding structure 1-2. The seal teeth 1-1 and the anti-rotation and flow guiding structure 1-2 are arranged on the inner surface of the seal stator 1. Along the flow direction, the anti-rotation and flow guiding structure 1-2 and multiple seal teeth 1-1 are arranged in sequence. That is, the anti-rotation and flow guiding structure 1-2 is installed at the inlet of the seal stator 1.

[0036] At the same time, multiple rotor bosses 2-1 are distributed on the rotor 2. The rotor bosses 2-1 are staggered between the seal teeth 1-1, which can divide a single seal chamber into two seal chambers, forming stronger vortex disturbances, thereby reducing the circumferential swirl and reducing the airflow excitation.

[0037] In some embodiments of the present invention, there is a radial radius gap C1 between the seal teeth 1-1 and the outer surface of the rotor 2. Its range can be determined by the calculation formula C1=(0.1-1)d / 1000, where d is the diameter of the rotor 2.

[0038] In some embodiments of the present invention, the height of the seal teeth 1-1 is H1, and its range is H1=[(d1-d) / 2]-C1, where d1 is the diameter of the seal stator 1 and C1 is the radial radius gap between the seal teeth 1-1 and the surface of the rotor 2.

[0039] In some embodiments of the present invention, the axial gap between adjacent seal teeth 1-1 is L1. In order to better generate vortices and improve the throttling effect of the seal teeth, it is necessary to ensure that the ratio of the volume to the gap between adjacent seal teeth is large enough (meeting the design requirements). It is defined that L1=(8-36)C1. Multiple circumferentially distributed rotor bosses 2-1 are arranged on the outer surface of the rotor 2. When setting the rotor bosses 2-1, the height H1 of the seal teeth 1-1 needs to be changed to ensure consistent seal gaps.

[0040] In some embodiments of the present invention, the anti-rotation and flow guiding structures 1-2 are evenly distributed circumferentially on the inner surface of the seal stator 1, and the number thereof is 60-150.

[0041] In some embodiments of the present invention, the radial radius gap distance between the anti-rotation and flow guiding structure 1-2 and the surface of the rotor 2 is C. To ensure the sealing performance of the steam seal and avoid rubbing between the anti-rotation and flow guiding structure 1-2 and the rotor 2, it is defined that the height H of the anti-rotation and flow guiding structure 1-2 is the same as the height H1 of the seal teeth 1-1, so that C = C1.

[0042] In some embodiments of the present invention, to ensure the flow guiding and throttling effects of the anti-rotation and flow guiding structure 1-2, the axial dimension of the anti-rotation and flow guiding structure 1-2 is B, and B=(1-3)L1. In addition, the axial gap between the anti-rotation and flow guiding structure 1-2 and the adjacent seal teeth 1-1 is L, and L=(1 / 5-1 / 2)L1.

[0043] In some embodiments of the present invention, the anti-rotation and flow-guiding structure 1-2 adopts a spiral-type anti-rotation and flow-guiding structure. The polar coordinate form of the logarithmic spiral is: r = aexp(θcotβ), where r is the polar radius; a is the starting polar radius, β is the spiral angle, and its value is the angle formed by the tangent at any point on the spiral and the line connecting the center of the circle to that point; θ is the circumferential angle between the logarithmic spiral at the polar radius and the starting point. The parametric coordinate form is: where b is the exponential factor. The starting polar radius a is the axial dimension of the sealing stator 1, and the value range of the exponential factor b is 0.25 to 0.4. Compared with the linear flow-guiding structure, the spiral flow-guiding structure can form a fluid convergence or divergence region; specifically, when the inlet region of the spiral flow-guiding structure is in millimeters, it causes a static pressure effect on the fluid, and the fluid flows directionally along the spiral direction of the spiral flow-guiding structure; when the inlet region of the spiral flow-guiding structure is in micrometers, the fluid is mainly affected by the dynamic pressure effect, and according to the spiral direction of the spiral flow-guiding structure, the fluid generates a pressure rise or drop to achieve the pressure regulation effect.

[0044] Reference Figure 2 and Figure 3 In some embodiments of the present invention, the spiral-type anti-rotation and flow-guiding structure is composed of a plurality of spiral flow-guiding plates 1-3 distributed circumferentially. It is installed at the inlet of the sealing stator 1 and is evenly distributed circumferentially along the inner surface of the sealing stator 1, and the number thereof is 60 - 150.

[0045] In some embodiments of the present invention, the height of the spiral flow-guiding plate 1-3 is the height H of the anti-rotation and flow-guiding structure 1-2, which is the same as the height H1 of the sealing tooth 1-1. The calculation formula is H = H1 = [(d1 - d) / 2] - C1, where d1 is the diameter of the sealing stator 1, and C1 is the radial clearance between the sealing tooth 1-1 and the surface of the rotor 2. Its axial dimension and the axial clearance with the adjacent sealing tooth are the same as those of the anti-rotation and flow-guiding structure 1-2, and the structure adopts the logarithmic spiral of the spiral-type anti-rotation and flow-guiding structure. That is, the calculation formula for its circumferential dimension B is B = (1 - 3)L1, where L1 is the distance between adjacent sealing teeth 1-1. Its axial distance L from the adjacent sealing tooth 1-1 is calculated as L = (1 / 5 - 1 / 2)L1.

[0046] In some embodiments of the present invention, the height of the spiral flow-guiding plate 1-3 is in millimeters. By setting the spiral flow-guiding plate 1-3, a millimeter-level spiral deep groove region is formed at the inlet of the sealing stator 1 to meet the conditions of the static pressure effect, where the static pressure effect of the fluid plays a major role; and the spiral direction of the spiral flow-guiding plate 1-3 is opposite to the rotation direction of the rotor 2, so that the circumferential swirling flow on the surface of the rotor 2 flows towards the inlet, reducing the gas flow excitation. In addition, part of the inlet fluid inflow is offset, reducing the leakage amount.

[0047] Reference Figure 4 and Figure 5, in some embodiments of the present invention, the anti-rotation and diversion structure 1-2 is a boss with spiral diversion grooves, which is composed of a stator boss 1-5 and spiral diversion grooves 1-6. The height H of the stator boss 1-5 is the same as the height of the sealing teeth, and its calculation formula is H1=(d1-d) / 2-C1, where d1 is the diameter of the sealing stator 1 and C1 is the radial radius clearance between the sealing teeth 1-1 and the surface of the rotor 2; the calculation formula for the axial dimension B of the stator boss 1-5 is B = 1-3L1, where L1 is the distance between adjacent sealing teeth 1-1.

[0048] In some embodiments of the present invention, the anti-rotation and diversion structure 1-2 is a boss with spiral diversion grooves, which is composed of a stator boss 1-5 and spiral diversion grooves 1-6. The stator boss 1-5 is arranged circumferentially, and the spiral diversion grooves 1-6 are evenly distributed circumferentially on the inner surface of the stator boss 1-5. The height of the stator boss 1-5 is the height H of the anti-rotation and diversion structure 1-2. The axial distance L between the stator boss 1-5 and the adjacent sealing teeth 1-1 is calculated by the formula L=(1 / 5-1 / 2)L1, where L1 is the distance between adjacent sealing teeth 1-1.

[0049] In some embodiments of the present invention, the spiral diversion grooves 1-6 are evenly distributed on the inner surface of the stator boss 1-5, and the number is 60-150.

[0050] In some embodiments of the present invention, the height of the stator boss 1-5 is in millimeters, and the spiral diversion grooves 1-6 are micron-level shallow grooves, that is, the height is in microns. Thus, a micron-level spiral shallow groove area is formed at the inlet to form a hydrodynamic pressure effect. The groove depth h of the spiral diversion grooves 1-6 is 0.001-0.01 times the radial radius clearance C1 to meet the formation conditions of the main hydrodynamic pressure effect of the fluid. The specific groove depth needs to be solved and calculated in combination with the fluid continuity equation, momentum equation and energy equation.

[0051] In some embodiments of the present invention, the helix direction of the spiral diversion grooves 1-6 can meet the throttling or diversion function according to actual needs. Specifically: when the helix direction of the spiral diversion grooves 1-6 is opposite to the rotation direction of the rotor 2, the air flow on the inlet side generates a pressure rise along the spiral diversion grooves 1-6, so as to form a high-pressure sealing area at the end of the spiral diversion grooves 1-6, achieving a throttling effect; when the helix direction of the spiral diversion grooves 1-6 is the same as the rotation direction of the rotor 2, the air flow on the inlet side generates a pressure drop along the spiral diversion grooves 1-6, reducing the impact vibration caused by the high-pressure air flow, achieving a diversion effect and improving the sealing performance.

[0052] The following are two specific embodiments of the present invention.

[0053] Embodiment 1:

[0054] Aiming at the goals of stable sealing and low leakage under high-pressure conditions, a radial seal integrating anti-rotation and diversion based on the hydrostatic effect is proposed. The following is a further detailed description in conjunction with the accompanying drawings:

[0055] Referring to Figure 2 and Figure 3 , the present invention includes a sealing stator 1, sealing teeth 1-1, and a spiral guide vane 1-3, where:

[0056] The sealing teeth 1-1 and the spiral guide vane 1-3 are installed on the inner surface of the sealing stator 1 and have a radial clearance with the rotor 2. In this embodiment, the diameter d1 of the sealing stator 1 is 1101 mm, and the diameter d of the rotor 2 is 1088.2 mm.

[0057] The radial clearance C1 between the outer surface of the sealing teeth 1-1 and the rotor 2, the range calculation formula of which is C1 = (0.1 - 1)d / 1000. Then, the height H of the sealing teeth 1-1 is determined, and the calculation formula is H = (d1 - d) / 2 - C1, where d is the diameter of the rotor 2 and d1 is the diameter of the sealing stator 1. The range of the radial clearance C1 is 0.1088 - 1.088 mm, and the range of the height H of the sealing teeth 1-1 is 5.312 - 6.291 mm. In this embodiment, the radial clearance C1 is 0.6 mm, and the height H of the sealing teeth 1-1 is 5.8 mm.

[0058] The axial clearance between adjacent sealing teeth 1-1 is L1. In order to better generate eddy fluid and improve the throttling effect of the sealing teeth, it is necessary to ensure that the ratio of the volume between adjacent sealing teeth to the clearance is large enough. It is defined that L1 = 8 - 36C1, and its range is 4.8 - 21.6 mm. In this embodiment, it is taken as 10 mm.

[0059] In this embodiment, the anti-rotation and diversion structure 1-2 adopts a spiral guide vane 1-3. The polar coordinate form of the logarithmic spiral is: r = r0exp(θtanβ), where r is the polar radius; r0 is the starting polar radius; β is the spiral angle, and its value is the angle formed by the tangent at any point on the spiral and the line connecting the center of the circle to that point; θ is the circumferential angle between the logarithmic spiral at the polar radius and the starting point. The parametric coordinate form is: where a is the starting polar radius and b is the exponential factor. In this embodiment, the starting polar radius a is taken as 36.6 mm, and the exponential factor b is taken as 0.4.

[0060] The spiral guide vane 1-3 is located at the inlet of the sealing stator 1, and it is evenly distributed circumferentially along the inner surface of the sealing stator 1, and the number is 72.

[0061] The radial radius clearance between the spiral deflector 1-3 and the surface of the rotor 2 is C. To ensure the sealing performance of the steam seal and avoid rubbing between the spiral deflector 1-3 and the rotor 2, it is defined that the height H of the spiral deflector 1-3 is the same as the height H1 of the sealing tooth 1-1, both being 5.8 mm, so that C = C1 = 0.6 mm.

[0062] The axial clearance between the spiral deflector 1-3 and the adjacent sealing tooth 1-1 is L. To ensure the influence of the fluid flow pattern and pressure change caused by the anti-rotation deflector on the fluid at the sealing tooth, it is defined that L = 1 / 5 - 1 / 2L1, and its range is 2 - 5 mm. In this embodiment, 2 mm is taken.

[0063] The axial dimension of the spiral deflector 1-3 is B. To ensure the flow guiding or throttling effect brought by the deflector, it is defined that B = 1 - 3L1, and its range is 10 - 30 mm. In this embodiment, 20 mm is taken.

[0064] By constructing the spiral deflector 1-3, a spiral deep groove area with a depth of 5.8 mm is formed at the inlet. Here, the hydrostatic pressure effect of the fluid plays a major role. Since its rotation direction is opposite to the rotation direction of the rotor, the circumferential swirl flow flows towards the inlet, reducing the gas flow excitation and the leakage amount.

[0065] Embodiment 2:

[0066] For different sealing inlet pressures, an anti-rotation and flow guiding integrated radial seal based on the hydrodynamic effect is proposed. The following is a further detailed description with reference to the accompanying drawings:

[0067] Refer to Figure 4 and Figure 5 , the present invention includes a sealing stator 1, a sealing tooth 1-1, a stator boss 1-5 and a spiral flow guiding groove 1-6, wherein:

[0068] The stator boss 1-5 and the sealing tooth 1-1 are located on the inner surface of the sealing stator 1, and spiral flow guiding grooves 1-6 are evenly opened circumferentially on the inner surface of the stator boss 1-5, and the number is 72.

[0069] In this embodiment, the dimensions of each part are the same as those in Embodiment 1. Only the anti-rotation and flow guiding structure 1-2 adopts a boss with a spiral flow guiding groove, which is composed of the stator boss 1-5 and the spiral flow guiding groove 1-6.

[0070] The radial radius clearance between the stator boss 1-5 and the surface of the rotor 2 is C. To ensure the sealing performance of the steam seal and avoid rubbing between the stator boss 1-5 and the rotor 2, it is defined that the height H of the stator boss 1-5 is the same as the height H1 of the sealing tooth 1-1, both being 5.8 mm, so that C = C1 = 0.6 mm.

[0071] The axial gap between the stator boss 1-5 and the adjacent sealing tooth 1-1 is L. To ensure the influence of the fluid flow pattern and pressure change caused by the spiral guide groove 1-6 on the fluid at the sealing tooth, it is defined that L = 1 / 5 - 1 / 2L1, and its range is 2 - 5 mm. In this embodiment, 2 mm is taken.

[0072] In this embodiment, the stator boss 1-5 is arranged at the inlet of the inner surface of the stator, and its axial dimension is B. To ensure the guiding or throttling effect brought by the spiral guide groove 1-3, it is defined that B = 1 - 3L1, and its range is 10 - 30 mm. In this embodiment, 20 mm is taken.

[0073] The groove type of the spiral guide groove 1-6 is a logarithmic spiral line. The polar coordinate form of the logarithmic spiral line is: r = r0exp(θtanβ), where r is the polar radius; r0 is the starting polar radius; β is the spiral angle, and its value is the angle formed by the tangent at any point on the spiral line and the line connecting the center of the circle to that point; θ is the circumferential angle between the logarithmic spiral line at the polar radius and the starting point. The parametric coordinate form is: where a is the starting polar radius and b is the exponential factor. In this embodiment, the starting polar radius a is taken as 36.6 mm and the exponential factor b is taken as 0.4.

[0074] The depth h of the spiral guide groove 1-6 is 0.001 - 0.01 times the radial radius gap C1, and its range is 0.6 - 6 μm. In this embodiment, 1 μm is taken.

[0075] The spiral guide groove 1-6 is a micron-level shallow groove, and the fluid is mainly affected by the hydrodynamic pressure effect. According to the fluid continuity equation, momentum equation, and energy equation, when the sealing inlet pressure is relatively low, it can be obtained that the rotor rotation direction is opposite to the spiral direction of the spiral guide groove, and the inlet-side air flow generates a pressure rise along the spiral groove, thereby forming a high-pressure sealing area at the end of the spiral groove to achieve a throttling effect.

[0076] The spiral guide groove 1-6 is a micron-level shallow groove, and the fluid is mainly affected by the hydrodynamic pressure effect. According to the fluid continuity equation, momentum equation, and energy equation, when the sealing inlet pressure is relatively high, it can be obtained that the rotor rotation direction is the same as the spiral direction of the spiral guide groove, and the inlet-side air flow generates a pressure drop along the spiral groove, reducing the impact vibration caused by the high-pressure air flow, achieving a guiding effect, and improving the sealing performance.

Claims

1. An anti-rotation and flow-guiding integrated radial seal based on dynamic and static pressure effects, characterized in that: It comprises an anti-rotation flow guide structure (1-2) and a plurality of sealing teeth (1-1) which are sequentially arranged on the inner surface of a sealing stator (1) along the flow direction, wherein the anti-rotation flow guide structure (1-2) is installed at the inlet of the sealing stator (1).

2. The anti-rotation and flow-guiding integrated radial seal based on dynamic and static pressure effects according to claim 1 is characterized in that: There is a radial radius gap C1 between the sealing tooth (1-1) and the surface of the rotor (2), which is calculated by: C1 = (0.1 ~ 1) d / 1000, and the calculation formula of H1 of the sealing tooth (1-1) is H1 = [(d1-d) / 2]-C1, where d1 is the diameter of the sealing stator (1), d is the diameter of the rotor (2), and the calculation formula of the axial gap L1 between adjacent sealing teeth (1-1) is L1 = (8 ~ 36) C1.

3. The anti-rotation and flow-guiding integrated radial seal based on dynamic and static pressure effects according to claim 2 is characterized in that: The height H of the anti-rotation flow guide structure (1-2) is H1, so that C=C1, wherein C is the radial radius gap between the anti-rotation flow guide structure (1-2) and the surface of the rotor (2); the axial gap between the anti-rotation flow guide structure (1-2) and the adjacent sealing tooth (1-1) is L, L=(1 / 5 to 1 / 2)L1; the axial dimension of the anti-rotation flow guide structure (1-2) is B, B=(1 to 3)L1.

4. The anti-rotation and flow-guiding integrated radial seal based on dynamic and static pressure effects according to claim 1, 2 or 3, characterized in that: The anti-rotation flow guiding structure (1-2) is a spiral anti-rotation flow guiding structure.

5. The anti-rotation and flow-guiding integrated radial seal based on dynamic and static pressure effects according to claim 4 is characterized in that: The polar coordinate form of the logarithmic spiral of the spiral anti-rotation guide structure is: r = aexp (θcotβ), wherein r is the polar diameter; a is the starting polar diameter, i.e., the axial dimension of the sealing stator (1); β is the spiral angle, which is the angle formed by the tangent at any point of the spiral and the line connecting the center of the circle to the point; θ is the circumferential angle between the logarithmic spiral at the polar diameter and the starting point. The parameter coordinate form is: Where b is an exponential factor, ranging from 0.25 to 0.

4.

6. The anti-rotation and flow-guiding integrated radial seal based on dynamic and static pressure effects according to claim 4 is characterized in that: The spiral anti-rotation flow guide structure is composed of a plurality of spiral flow guide plates (1-3) distributed along the circumferential direction, the height of the spiral flow guide plates (1-3) being the same as the height H of the anti-rotation flow guide structure (1-2), the axial dimension of the spiral flow guide plates (1-3) and the axial clearance with the adjacent sealing teeth being consistent with those of the anti-rotation flow guide structure (1-2), and the structure adopts the logarithmic spiral of the spiral anti-rotation flow guide structure.

7. The anti-rotation and flow-guiding integrated radial seal based on dynamic and static pressure effects according to claim 6 is characterized in that: The height of the spiral guide plate (1-3) is in the millimeter range, thereby forming a millimeter-level spiral deep groove area at the inlet of the sealed stator (1), satisfying the static pressure effect condition; and the rotation direction of the spiral guide plate (1-3) is opposite to the rotation direction of the rotor (2), so that the circumferential swirl flow on the surface of the rotor (2) flows toward the inlet.

8. The anti-rotation and flow-guiding integrated radial seal based on dynamic and static pressure effects according to claim 4 is characterized in that: The spiral anti-rotation flow guide structure is a boss with a spiral flow guide groove, which is composed of a stator boss (1-5) along the circumferential direction and a plurality of spiral flow guide grooves (1-6) evenly distributed along the circumferential direction on the inner surface of the stator boss (1-5), and the height of the stator boss (1-5) is the height H of the anti-rotation flow guide structure (1-2).

9. The anti-rotation and flow-guiding integrated radial seal based on dynamic and static pressure effects according to claim 8 is characterized in that: The height of the stator boss (1-5) is in the millimeter order, and a spiral guide groove (1-6) with a height of the micrometer order is provided on its inner surface, thereby forming a micrometer-order spiral shallow groove area at the inlet to form a dynamic pressure effect. The groove depth of the spiral guide groove (1-6) is h=(0.001-0.01)C1, so as to meet the conditions for the fluid to be mainly subjected to the dynamic pressure effect, wherein C1 is the radial radius gap between the sealing tooth (1-1) and the surface of the rotor (2).

10. The anti-rotation and flow-guiding integrated radial seal based on dynamic and static pressure effects according to claim 8, characterized in that: When the rotation direction of the rotor (2) is opposite to the rotation direction of the spiral guide groove (1-6), the airflow on the inlet side generates a pressure rise along the spiral guide groove (1-6), thereby forming a high-pressure sealing area at the end of the spiral guide groove (1-6), thereby achieving a throttling effect; when the rotation direction of the rotor (2) is the same as the rotation direction of the spiral guide groove (1-6), the airflow on the inlet side generates a pressure drop along the spiral guide groove (1-6), thereby reducing the impact vibration caused by the high-pressure airflow and achieving a guide effect.

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