CN100557197C

CN100557197C - A kind of mixed flow type turbine vane

Info

Publication number: CN100557197C
Application number: CNB2006100255063A
Authority: CN
Inventors: 孙敏超; 孙正柱
Original assignee: Individual
Current assignee: Individual
Priority date: 2006-04-07
Filing date: 2006-04-07
Publication date: 2009-11-04
Anticipated expiration: 2026-04-07
Also published as: CN101050710A

Abstract

The invention discloses a kind of combined flow turbine semi-open type (or enclosed) impeller that turbo-expander is used in internal combustion engine turbocharger, middle-size and small-size gas turbine installation, chemical industry and refrigeration (gas liquefaction with the separate) equipment.By flow field analysis, the present invention has illustrated wheel rotation angular velocity , the interior meridional stream line slope inclination angle δ of wheel and meridional stream line is to flowing in taking turns the operating mode factor and the geometrical factor of material impact to be arranged at the semidiameter Δ r of impeller inlet/outlet.Also corrected simultaneously and spread the cacodoxy of deep and broad " work done of Coriolis power " and the impeller design misdirection principle of corresponding association in the prior art.On this basis, the present invention has set up the selection that obtains high efficiency mixed flow type turbine vane meridian profile construction shape important geometric parameter under the higher specific speed operating mode and has recommended the scope of using.

Description

Mixed-flow turbine impeller

Technical Field

The invention relates to a turbine wheel for a turbo-expander in internal combustion engine turbochargers, medium and small gas turbine units, chemical and refrigeration (gas liquefaction and separation) equipment.

Background

According to the direction of fluid flow in the turbine impeller, the turbine can be divided into three types, namely an axial flow turbine, a radial flow turbine and a mixed flow turbine. Wherein, the axial flow turbine, the fluid flows through along the direction approximately parallel to the rotating shaft of the impeller; the radial turbine, the fluid flows into the impeller from the rim of the impeller to the axial direction of the rotating shaft along the radial direction approximately vertical to the rotating shaft of the impeller, and turns into axial outflow at the outlet of the impeller; a mixed-flow turbine (also called a diagonal-flow turbine) is an intermediate form between an axial-flow turbine and a radial-flow turbine, in which fluid flows through an impeller along a conical surface inclined to an impeller rotation axis. Mixed flow turbines have been widely used and developed in recent years as an improvement in radial turbines to achieve high efficiency at high specific speeds. In fact, the inclination of the inlet and outlet edges of the mixed-flow turbine impeller blade (increasing the height of the blade in the through-flow part) and the increase of the diameter of the outlet edge are the main geometrical characteristic changes of the impeller shape, which are the natural trends of the blade shape of the radial-flow turbine impeller to adapt to the flow rate development towards large capacity under the condition of high rotating speed. The axial length of the through-flow part of the mixed-flow turbine impeller is properly increased compared with that of the radial-flow turbine impeller, so that the flow field in the impeller is improved, and the smoothness of the flow is increased.

The most obvious advantage of mixed flow turbines is that they can achieve high efficiency at high specific speeds compared to radial turbines. Due to the important position of the turbine efficiency in the research of the turbine performance, the achievement of higher efficiency at high specific speed is also the most important research direction in the mixed flow turbine.

In the practice of applying mixed-flow turbine to turbocharger (such as RR151 turbocharger of ABB company, Garrett turbocharger of Honeywell company, KTR150 turbocharger manufactured by komatsu, RH-3 turbocharger of shichuan island company, etc.), the highest isentropic efficiency of mixed-flow turbine is proved to be higher than that of radial-flow turbine, and the highest isentropic efficiency is improved by about 5% on average. However, the theoretical studies of the radial turbine and the mixed-flow turbine are very insufficient. Even why is mixed-flow turbine more efficient than radial-flow turbine? What are factors that limit the improvement in radial turbine efficiency? How to design a mixed-flow turbine to ensure high efficiency? The prior art is still in the exploration stage, wherein errors in basic concepts and concepts affect correct recognition, analysis and judgment of the affairs, so that correct and effective guidance on the design cannot be provided.

For the 'simple mechanism' that the efficiency of the mixed-flow turbine is better than that of the radial turbine, the prior art is attributed to the geometrical characteristic that the blade inlet edge of the mixed-flow turbine wheel is inclined (inclined by an angle theta with the rotating shaft of the turbine). It is considered that this configuration is advantageous for forming a "forward-curved blade" shape at a portion near the hub along the oblique edge of the blade inlet, thereby reducing "incident loss" of the air flow and "turning loss" generated when the fluid flows axially in the impeller flow passage from a radial direction, so that the mixed-flow turbine can greatly improve the efficiency as compared with the radial-flow type (see, for example, H145 turbocharger using mixed-flow turbine and its application to Z6170 diesel engine, 6 th and 19 th to 20 th pages in 2003, and related comments and reports on "development of mixed-flow turbine of turbocharger for vehicle, diesel engine, 2000, 6 th and 14 th to 18 th pages in japan, ltd. In fact, in some radial-flow turbine impellers of the six and seventies of the last century, structures in which the blade inlet edges are inclined (such as the 4HD turbocharger of Schwizer, the NR turbocharger of MAN, the ZY-120 turbocharger of Chongqing heavy-duty automobile institute, and the ZY-120 turbocharger of Chongqing automobile engine plant) were adopted, but the turbine stages did not show a significant efficiency improvement advantage over the turbine stages having conventional radial-flow turbine impellers in which the blade inlet edges were parallel to the rotation axis. In addition, for a conventional radial turbine impeller, the distribution types of different blade inlet geometric angles such as forward bending, radial bending, backward bending and forward bending and backward sweeping can be respectively formed from the hub to the rim on the non-inclined blade inlet edge through the blade modeling. The forward-curved backward-swept impeller has the same blade inlet geometric angle distribution type as that required by the mixed-flow turbine impeller along the inclined inlet edge of the blade (as shown in Chinese patent publication No. CN01231703.9 mixed-flow turbine impeller (FIG. 3)). The comparative analysis of the quasi-ternary flow field of the conventional radial-flow turbine impeller with different blade inlet geometric angle distribution types shows that the turbine impeller with the forward-bent and backward-swept blade inlet geometric angle distribution really has better variable working condition adaptability, can reduce the flow loss of the change and proliferation of different inflow attack angles, and improves the turbine efficiency under the variable working condition, but does not obviously improve the highest isentropic efficiency value of the turbine grade (see the influence of the centripetal turbine impeller inlet profile on the variable working condition performance, such as creep peak, the dynamic machinery and engineering thermophysics, national dynamic machinery and engineering thermophysics, the report of the youth academic paper of the national dynamic machinery and engineering thermophysics university, the collection of literature, the xi ' an ' xi ' ann transport university press, 10 months in 1989, and 659-663 pages). Thus, the inclination of the blade inlet edge of the turbine wheel and the forward-curved swept-back type blade inlet geometry angular distribution along the blade inlet edge are not the unique geometric features of mixed-flow turbine wheels, as are radial-flow turbine wheels. Therefore, they are not the essential reason for the mixed flow turbine efficiency to be significantly higher than the radial turbine efficiency. In this regard, the results of turbocharger to diesel matching experiments also corroborate this view: by changing the inclination angle of the inlet edge of the turbine blade and the geometric angle distribution of the inlet edge of the turbine blade, obvious efficiency benefits cannot be obtained, namely the lowest fuel consumption of the diesel engine is not greatly reduced.

Disclosure of Invention

The invention aims to determine the factors which have the biggest influence on the flow of the axial flow turbine impeller, the radial flow turbine impeller and the mixed flow turbine impeller and are different from each other among a plurality of geometrical factors and working condition factors which influence the flow by research, analysis and comparison of the flow in the axial flow turbine impeller, the radial flow turbine impeller and the mixed flow turbine impeller; the method corrects the profound and broad wrong understanding and concept in the prior art; based on the structure, a reasonable geometric parameter recommended range for obtaining the meridian section structural shape of the high-efficiency mixed flow turbine impeller is established.

In view of the flows in axial, radial and mixed flow turbines, can be approximately simplified into a rimA combination of blade cascade flows of arbitrary gyration planes. The geometrical shape of the family of gyration surfaces is formed by the rotation of a family of meridian flow lines in the impeller around the same rotation shaft (turbine shaft). Therefore, the flow surface of the axial flow turbine is a group of gyration surfaces with the shape similar to the shape of a coaxial cylindrical surface; the flow surface of the radial turbine is a group of gyration surfaces which are approximately radial planes at the inlet of the impeller and approximately cylindrical surfaces at the outlet; the flow surface of the mixed flow turbine is a gyration surface (shown in figure 1) with a group of impellers with an inlet approximately in a conical surface and an outlet approximately in a cylindrical surface. Clearly, the divergence in the geometry of the flow surfaces (i.e., the meridional flow lines) is the first important geometric factor that produces significant divergence in the flows in axial, radial, and mixed flow turbine stages. The influence of the meridian flow shape on the vane cascade flow field of the gyration surface is mainly reflected in the distribution delta (r) or delta (z) of the slope inclination angle delta (arctg (dr/dz) value along the meridian flow line and the radius difference delta r (r) of the meridian flow line at the inlet and the outlet of the impeller₁-r₂) These two geometric elements (fig. 2).

According to the blade cascade flow theory of any gyration surface in the mechanical fluid dynamics of the impeller, the steady flow of the non-viscous fluid circumfluence blade cascade must satisfy the absolute motion non-rotation equation

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When the fluid flows along any surface of revolution, the degree of rotation

Normal to the plane of revolution only

Has a component ofThen the above equation is

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That is to say, the motion equation followed by the fluid flowing along any gyration surface

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In the formula-fluid motion velocity (absolute velocity) measured by an absolute coordinate system fixed with the stationary part of the turbine stage (housing, nozzle ring);-fluid movement speed (relative speed) measured by a relative coordinate system fixed with the rotating impeller of the turbine stage;

angular speed of rotation of the impeller (attached)Fig. 3).

From this equation of motion: the fluid motion in the gyration surface blade cascade in the rotary impeller is a relative motion and swirl flow, and the swirl value is-2 omega sin delta. When δ is 0 (axial flow turbine, flow surface is cylindrical surface), the fluid flows throughThe flow of any one of the cylindrical surface cascades in the axial-flow turbine impeller rotating at the angular velocity is completely the same as the flow of the same cylindrical surface cascades when the cascades are static and do not rotate (omega is 0), that is, the flow in the axial-flow turbine impeller is not influenced by the rotating speed (the angular velocity)) The influence of (c). This is the essential reason why the flow in an axial turbine wheel is distinguished from the fluid movement in a radial turbine wheel and a mixed-flow turbine wheel. For the flow near the inlet of the radial turbine impeller, the flow field in the part of the area is subjected to the maximum influence of the rotation generated by the rotation angular velocity due to the fact that delta is approximately equal to 90 degrees. In contrast, the flow near the mixed flow turbine wheel inlet, due to 0 < δ < 90 °, affects the rotation due to angular velocity between the axial and radial turbines. As for the axial outlet portion of the radial turbine and mixed-flow turbine wheel, the flow on the flow surface is substantially unaffected by the angular velocity, as with the axial turbine wheel, due to δ ≈ 0 of the flow surface of this portion.

The velocity distribution of the flow field in the blade grid of any turning surface in the impeller can be clearly shown by the following method: in practice, the rotational speed is

The flow of the blade cascade with any circumgyration surface of the circumgyration flow under the working condition of G flow rate can be decomposed into the flow of the same circumgyration surface blade cascade with the flow rate of G and the flow of the same circumgyration surface blade cascade with the flow rate of G being static and not rotating (omega is 0) and the flow of the other flow rate of zero (G is 0, the inlet and the outlet of the blade cascade flow passage are respectively sealed along the circumference), the flow direction and the inlet and the outlet of the blade cascade flow passage are respectively sealed along the circumference

The rotation directions determined by the component vectors in the normal direction of the rotation surface are opposite, and the variable strength circulation with the vortex strength of 2 omega sin delta flows in the same rotation surface cascade flow channel for superposition (shown in figure 4). Then, the flow rate is G and the rotation speed is G

In the flow of the blade cascade of any circumgyration surface of the circumambient flow under the working condition, the velocity vector of any point in the blade cascade is accurately equal to the sum of the velocity vectors of the same point in the two flow fields. Because of the circumfluence (circumvortex) in the closed flow passage turning and

the component in the normal direction of the gyration surface turns reversely, and after the component is superposed with the through-flow, the flow velocity near the suction surface of the cascade runner is increased, and the flow velocity near the pressure surface is reduced. The greater the value of the vorticity 2 ω sin δ, the greater the difference in velocity between the pressure and suction surfaces it causes (fig. 5). The velocity gradient is one of the main power sources of the low kinetic energy fluid in the boundary layer of the hub, which makes the transverse migration motion (secondary flow) from the pressure surface to the suction surface along the hub surface, and the magnitude of the end loss is greatly influenced.

In summary, the influence of the inclination angle δ of the meridian streamline slope on the flow of the blade cascade of the internal revolution surface of the wheel is very broad, and far from the influence of the difference of the size of the turning loss caused by different turning slowness degrees of the flow of the meridian plane in the impeller recognized by the prior art, the influence of the rotation speed ω applied to the flow at each point in the flow field of the blade cascade of the internal revolution surface of the whole wheel in the form of the rotation 2 ω sin δ is applied. In addition, in many "mixed flow turbine" designs of the prior art, although the inlet edge of the impeller blade is already made to be inclined (at an angle θ with the rotating shaft), the flow direction of the meridional velocity which is divided into two equal parts of the "average flow velocity" according to the step flow and flows out from the outlet of the nozzle ring blade (or the single-channel or double-channel vaneless volute) is not perpendicular to the inclined inlet edge of the impeller, and the design of the radial turbine is still adoptedThe direction delta ≈ 90 ° flows into the impeller of the "mixed flow turbine". For this "mixed flow turbine", the substance is still a radial flow turbine, and the flow field in the wheel does not change significantly. Thus, a correct, high efficiency mixed flow turbine stage design, which is not sufficient with only a view to the design of the impeller, must also be well matched to the design of the nozzle ring or the vaneless volute, i.e. the nozzle ring or the vaneless volute should be constructed with a meridian velocity direction which matches the mean velocity at its outlet with δ₁The ability to flow in the direction of the impeller of ≈ (90- θ).

As for the geometrical element of the second term of the meridian streamline shape, the radius difference Δ r ═ r (r)₁-r₂) The influence on the flow field of the blade cascade of the inner gyration surface of the wheel is also applied together with the rotating speed omega like the first geometric element delta. Each gyration surface (formed by different meridian flow lines in the wheel revolving around the same shaft axis) in the gyration surface family in the wheel has the radius difference delta r (r ═ r)₁-r₂) Although different from each other, the radius difference Δ r affects the flow on each of the planes of gyration in the same manner. In order to reflect the influence of the flow on the whole impeller, the flow is analyzed along a middle gyration surface (the flow between a hub gyration surface and a rim gyration surface is equally divided into two halves of the middle gyration surface which is formed by the gyration of a middle streamline around a rotating shaft in an impeller meridian flow family) as a representative. For the intermediate gyration surface, the radius difference Delta R is the geometric mean radius R of the impeller inlet₁＝[(R_1sh ²+R_1h ²)/2]^0.5Geometric mean radius R of impeller outlet₂＝[(R_2sh ²+R_2h ²)/2]^0.5In the difference between,. DELTA.R ═ R₁-R₂) (FIG. 6).

A common geometric feature of radial and mixed-flow turbine wheels, as opposed to axial turbine wheels, is the presence of a significant radius difference ar. It illustrates that there is radial displacement or radial flow through radial flow, mixed flow turbine wheels. Because the fluid is formed by R with larger radius in the impeller rotating at high speed₁R with smaller radius of flow direction₂So that the flow process is performedMust overcome the centrifugal inertia force

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Can be achieved (where dm is the fluid infinitesimal mass). The effect of centrifugal inertial forces on this flow process. Can be reflected by the work it does on the movement of the fluid, i.e. for a unit mass of fluid

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In the formula

-the drawing speed (peripheral speed),

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this is a well-known Euler equation (basic equation for impeller mechanics) L that determines the magnitude of the amount of work performed by a fluid flowing through an impeller_u＝(C₁ ²-C₂ ²)/2+(W₂ ²-W₁ ²)/2+(U₁ ²-U₂ ²) The third component of/2.

To date, in the monographs and textbooks of numerous influential turbomachinery principles, turbomachinery hydrodynamics and turbochargers at home and abroad, the term "U" is used₁ ²-U₂ ²) The understanding that the mechanical explanation of the item/2 has errors; considered as "Coriolis force" work (of magnitude U)₁ ²-U₂ ²) A part of (a); it is also considered that in the centripetal flow of fluid in the radial and mixed flow turbine wheel, the generation of "Coriolis force" is not accompanied by the turning of the flow or the accompanying frictional loss, and a part of the effective work of the Coriolis force is substantially not energy loss. Thereby establishing a design turbineError guiding principle of impeller: the radius size difference delta R between the inlet and the outlet of the impeller is preferably increased as much as possible (R)₁-R₂) The proportion of the amount of the work of the Coriolis force in the wheel periphery work determined by the Euler equation is made as large as possible (see M.H. Waff, pneumatic thermodynamics and flow in turbomachinery, Beijing, mechanical industry Press, 1984, 8 th month, pages 112-115; a, a book of "gas turbine and gas turbine installation" of Kirillofu, Beijing mechanical industry Press, 1959, pp 157 to 158 and another book of "principles of turbomachinery", Beijing mechanical industry Press, 1982, pp 6, 575 to 582; the chapter of Zhumeilin, the chapter of turbocharger principle, Beijing, national defense industry Press, 6 months in 1982, pages 276-278; ju Da Xin, related discussion in "turbo charger and turbo charger", Beijing, mechanical industry Press 1992, 11 Yue.179-181, pages 224).

The reason for the above mentioned erroneous concepts is that the above mentioned documents are erroneous in the recognition and application of the following basic mechanical concepts:

● relating Coriolis acceleration

Considering "Coriolis force" acting on a unit mass of fluid, two completely different concepts in mechanics, namely "force" and "acceleration", are mixed up into the same concept. Any force must have a property that can be directly measured by a dynamometer (see B · Γ · nieri tokyo doff, book of theoretical mechanics (huangnings) — beijing · people education press-1964, 8 months, page 242), "Coriolis force" cannot be directly measured by a dynamometer, so it is not force. In the flow of fluid through the rotating impeller flowpath, there is no Coriolis force, but there is a Coriolis inertial force. Coriolis inertial force acting on a unit mass of fluid

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● Coriolis inertial forces exist in a relative coordinate system (non-inertial system) fixed to the impeller for rotation about the turbine axis at an angular velocity ω, which acts as a relative motion of the fluid flowing through the impeller flowpath. Displacement due to constant and relative movement of the force

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Are perpendicular (measured by a relative coordinate system) so that their work is constant at zero

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Coriolis inertial forces are absent in an absolute coordinate system (inertial system) that is fixed to the stationary turbine stage parts (housing, nozzle ring). Therefore, it cannot do work for the motion process of the involvement displacement dl (Udt) ω xrdt (measured by the absolute coordinate system) of the fluid existing only in the absolute motion. The magnitude that relates the force and displacement present in different coordinate systems together as a product is not conceptual and meaningful in Newton's mechanics.

The ● Coriolis inertial force is not the true force due to object interaction. The real force differs from the inertial force in that: the presence of real forces (such as pressure, friction, attraction, electromagnetic force, elastic force, etc.) is not changed by the choice of coordinate system, the effect being the same regardless of which coordinate system is used to describe it; the inertial force varies depending on the choice of the coordinate system. Another important difference is that the real force is present as a reaction force and the inertial force is absent. Therefore, the Coriolis inertia force cannot directly generate a relationship between the reaction force and the reaction force with the blades of the turbine wheel, and the work amount is exchanged. In fact, an impeller machine is a machine that effects a change in the energy level of a fluid by the exchange of work through the interaction of the blades with the fluid. When fluid flows through the rotating impeller, it acts on the blade surfaces in the form of pressure, the resultant of which produces a moment (torque) on the turbine shaft and power as the impeller rotates.

In fact, the Euler equation can be derived by using the theorem of kinetic energy in the mechanical fluid dynamics of the impeller, and the component terms (U) are strictly proved₁ ²-U₂ ²) And/2 is the work value of the centrifugal inertia force.

The following continues to analyze the effect of the speed of rotation ω and the difference in radius Δ R on the fluid movement within the turbine wheel. According to the equation of energy (stagnation enthalpy is conserved along the same flow line) and

in the formula, W₁，W₂-the relative flow rates at the inlet and outlet of the impeller, respectively; (i)₁-i₂) -isentropic enthalpy difference of the fluid at the inlet and outlet of the impeller; psi-flow velocity loss coefficient in impeller. It shows the flow velocity W at the outlet of the impeller under the action of the centrifugal inertia force field (centrifugal force field)₂Following difference value (U)₁ ²-₂ ²) Is increased and decreased. With (U)₁ ²-U₂ ²) Increase of term, W₂Not only can be smaller than W₁(reduced diffusion flow) and can even be reduced to zero (no flow in the impeller). This situation, in turn, highlights the difference between radial and mixed flow turbines and axial turbines: in axial flow turbines, as long as there is an enthalpy difference (i) within the wheel₁-i₂Greater than 0), W does not appear in the impeller₂＜W₁Diffusion flow of (2); for radial and mixed flow turbines, the turbine must be used

(i_{1} - i_{2}) > (U_{1}^{2} - U_{2}^{2})

Then the process is carried out. In the turbine wheel, W₂＜W₁Otherwise, the flow in the direction of flow will be diffusion flow, which will thicken and separate the boundary layer, resulting in a sharp increase in flow losses. Due to the fact that

Therefore, in order to control the deceleration effect of the flow velocity in the wheel under the high rotating speed condition and the corresponding pressure increase degree in front of the turbine impeller, the most effective measure is to make the control diameter difference delta R smaller to limit (U) when designing the geometric shape of the impeller flow passage₁ ²-U₂ ²) The value is exceeded.

It is evident that a reduction in ar will substantially attenuate the effect of the centrifugal field in the wheel, thereby causing a drop in fluid pressure at the inlet of the turbine wheel and a corresponding drop in fluid pressure in front of the nozzle ring. This is particularly important in the case of turbochargers for internal combustion engines, since a drop in the turbine inlet pressure leads to a drop in the exhaust gas back pressure of the internal combustion engine, which reduces the power consumption of the piston to overcome the back pressure effect when discharging the exhaust gas, i.e. the fuel consumption of the internal combustion engine is reduced.

In order to fully compare the flow differences between axial, radial and mixed flow turbines, it is also necessary to analyze and compare the "secondary flows" within the turbines. "secondary flow" inside the turbine wheel "The low-kinetic energy fluid in the boundary layer on the hub gyration surface (the closed impeller also comprises a rim gyration surface) migrates and flows from the pressure surface of the runner to the suction surface across the runner along the hub surface (and the rim surface-the closed impeller); the migration flow of low kinetic energy fluid on the blade surface from the blade root (hub position) to the blade tip (rim position) along the blade surface; and the gap air leakage at the rim gyration surface of the semi-open impeller and the flow of the blade scraping the boundary layer at the matched gap part of the inner shell wall of the static turbine volute are totally three parts. The differences in the "secondary flows" within the axial, radial and mixed flow turbine impeller wheels are primarily reflected in the first portion. Low kinetic energy fluid (relative flow velocity W) attached to the revolving surface of impeller hub (rim) of radial-flow turbine and mixed-flow turbine_B0) and moves along the plane of revolution of the hub (rim) under the combined action of the differential pressure force, the component force of the centrifugal inertia force along the plane of the hub (rim) and the Coriolis inertia force. Component omega along the hub (rim) plane in view of centrifugal inertia force²r sin δ is generally larger than the pressure difference between the pressure surface and the suction surface of the flow passage (along the normal direction of the main flow line, from the pressure surface to the suction surface — the direction of the action of the pressure difference force) generated by the main flow motion, and the pressure difference dp along the main flow direction is rhod [ (ω [ (. omega. ])²r²-W²)/2](where ρ is the fluid density) is small, and Coriolis inertial force-

Is of small magnitude (due to W)_BSmall) and so the combined effect of the "secondary flow" of the hub (rim) surface is that the low kinetic energy portion of the fluid in the boundary layer of the hub (rim) surface migrates along the hub (rim) surface from the near pressure zone of smaller radius to the suction zone of larger radius. The flow direction of the secondary flow on the hub (edge) surface and the main flow are in an oblique and reverse direction, so that the secondary flow continuously receives the scouring of the reverse main flow in the migration flow process of the secondary flow along the hub (edge) surface and flows into the main flow, and the separation flow is not easy to occur to cause large flow loss. Secondly, the circumferential width of the flow channel of the hub surface gyration surface is greatly changed from the contraction of the impeller inlet to the contraction of the impeller outlet. This results in a large flow width convergence gradient of the flow channel cross section, an increased degree of acceleration of the main flow, and a pressure difference between the pressure side and the suction side of the flow channelThe sharp decrease from the impeller inlet to the outlet, the small flow area of the hub surface near the considerable area of the impeller outlet, all of which greatly limits the existence and development of the 'secondary flow' on the hub surface gyration surface. This is fundamentally different from the generation, development and influence of the "secondary flow" on the hub face (cylindrical surface) of an axial flow turbine blade. The 'secondary flow' on the hub surface of the axial flow turbine blade flows along the 'oblique forward direction' of the migration of the main flow, and the 'secondary flow' accumulates, thickens and separates the low-kinetic-energy fluid of the boundary layer of the hub surface along the main flow to the suction surface (the back surface of the blade) of the flow channel close to the outlet, so that 'end loss' with a large magnitude is caused.

It is well known that of the three types of turbine stages, the highest isentropic efficiency value that can be achieved by an axial turbine stage is highest and the second is lowest by a mixed flow turbine when the vane height of the flow-through portion is sufficiently long (i.e., the stage flow is sufficiently large). This is mainly due to the difference in the adverse effect of the rotational speed, while the "end losses" are due to the fact that they have a severe effect only on the flow losses of the elementary stages near the hub and the rim, when the proportion of their area of influence is still small, averaged over the length of the blade. The blade height is correspondingly reduced along with the reduction of the flow, the proportion of the end loss in the flow loss is correspondingly increased, but the end loss influence is sequenced to be the heaviest to the axial flow turbine, the mixed flow type is the second to the radial flow type, and the radial flow type is the lightest, so the three components are close to each other when the highest efficiency value is synchronously reduced. When the flow rate is reduced to a level where the blade height of the axial turbine is reduced to a level where the "secondary flow" influence area at the two ends of the blade root (hub surface) and the blade tip (rim) occupies a certain proportion of the entire blade height, the maximum efficiency value that can be achieved by the radial turbine exceeds that of the axial turbine. When the 'secondary flow' influence areas of the two ends of the axial-flow turbine blade root and the top are converged together, the highest isentropic efficiency value of the axial-flow turbine stage is sharply reduced, and the axial-flow turbine stage is not suitable for working under the working condition of small flow and is applied to radial and mixed-flow turbines instead.

From the viewpoint of increasing the blade height and reducing the end loss to improve the highest isentropic efficiency value of the turbine stage, it is advantageous to increase the axial width and the inclination angle theta value of the inlet edge of the blade to further increase the blade height of the inlet edge of the mixed flow turbine impeller. This measure also applies to the outlet edges of the nozzle ring vanes and the impeller (increasing the angle of inclination γ).

In summary, the impeller rotational speed (angular velocity)

) The method is the most important working condition factor for enabling axial flow, radial flow and mixed flow type turbine impeller flow fields to be different from each other. The higher the rotational speed, the greater the adverse effect of centrifugal inertial forces on the flow in the wheel. Compared with the structure of a radial-flow turbine impeller, the structure of the mixed-flow turbine impeller is more beneficial to weakening the action of centrifugal inertia force, so that the highest isentropic efficiency of the mixed-flow turbine stage is obviously higher than that of the radial-flow turbine stage in the reasonable design of effectively inhibiting the influence of the centrifugal inertia force.

For the same reason: for the comparison of the efficiency of radial turbine and mixed-flow turbine with similar wheel diameter size varying with the rotation speed (expansion ratio), why the isentropic efficiency difference between them is larger at high rotation speed than at low rotation speed? And can explain why are the maximum isentropic efficiency values achievable by axial, radial and mixed flow turbine stages at high flow but relatively low speed? This situation is particularly noticeable in recent years for applications to large and medium-sized turbochargers. In the turbine stage of the turbocharger applied to the large and medium power diesel engines, the same supercharging ratio (the speed U of the impeller wheel circumference) is generated due to large flow (the length of the blade and the diameter of the impeller wheel rim are large)₁＝ωR₁Square of the pressure is in direct proportion) is far lower than the angular velocity required by a small flow (small wheel diameter) impeller reaching the same pressure ratio, so that the efficiency of the radial and mixed flow turbine stages can be obviously improved and is close to that of the axial flow turbine. It is thus possible to have a part of the efficient mixed-flow and radial-flow turbines, which, because of their simplified structure and low manufacturing costs, have the advantage of not significantly reduced performance, come to a new trend in this field of application, which is originally occupied entirely by axial-flow turbines.

In the prior analysis, on the basis of comparing the influence of the geometric parameters of the impeller and the rotating speed of the impeller on the flow in the wheel, the invention establishes the reasonable recommended range of the important geometric parameters for obtaining the meridian section structural shape of the high-efficiency mixed-flow turbine impeller as follows:

R_2sh≤R_1sh；

＝R₂/R₁＝0.73-0.93；θ＝20°～70°；γ＝-15°～30°；

＝B/(2R₁)＝0.45～0.60；

＝l₁/(2R₁)＝0.16～0.25.

wherein,

geometric mean radius R of the outlet edge of the impeller blade₂Geometric mean radius R of inlet edge of impeller blade₁The ratio of (a) to (b). The value reflects the geometric mean radius difference DeltaR (R) of the inlet and outlet edges of the impeller blade₁-R₂) Relative impeller blade inlet edge geometric mean radius R₁The magnitude of (a) is greater than (b),

＝1-ΔR/R₁。

theta is the included angle between the inlet edge of the impeller blade and the axis of the rotating shaft. It approximately reflects the slope inclination angle delta of the meridian flow line in the wheel on the bevel edge of the blade inlet of the impeller₁Magnitude of ≈ (90- θ);

the axial length B of the flow-through portion of the impeller blades and the geometric mean diameter of the edge of the impeller blade inlet

D_{1_{m}} = 2 R_{1}

The ratio of (a) to (b). Increase of

The variation smoothness of the meridian flow slope inclination angle delta from the impeller inlet to the impeller outlet can be increased;

length l of the inclined edge of the impeller blade inlet₁Geometric mean diameter of impeller blade inlet edge

The ratio of (a) to (b). The magnitude of which is the relative height of the blade.

Gamma is the included angle between the outlet edge of the impeller blade and the radius direction. Increasing the angle gamma increases the length of the blade outlet side and increases the width of the blade flow channel outlet throat (increases the outlet area).

To date, the prior art has not introduced a systematic and complete recommended application range for the selection of geometric parameters of mixed-flow turbine impellers, and only reports the individual design and application examples. The mixed flow turbine is an improvement of the radial flow turbine under the condition of high specific speed, and the parameters of the mixed flow turbine are selected by taking the upper limit of the recommended range of the corresponding geometric parameters of the radial flow turbine and then appropriately amplifying and changing. The recommended range of geometric parameters of the prior art radial turbine impeller is as follows: r₁＝R_1sh＝R_1h；R_2sh/R₁＝0.7～0.86；R₂/R₁＝0.5～0.6；θ＝0°；

＝B/(2R₁)＝0.31～0.36；

＝l₁/(2R₁)＝0.08～0.15；γ＝0°～10°。

Comparing the present invention with the recommended range of the geometric parameters of the radial-flow turbine impeller, it can be seen that the geometric parameters of the mixed-flow turbine impeller recommended by the present invention are greatly increased compared with the recommended values of the radial-flow turbine in the prior art, showing the tendency of the shape of the mixed-flow turbine impeller to the shape of the axial-flow turbine impeller with high efficiency.

The mixed-flow turbine wheel designed in view of the invention has a hub outlet radius R_2hThe value is generally larger, and in order to recover the exhaust energy and reduce the residual speed loss, an exhaust diffuser (which is arranged at the outlet of a turbine impeller and connected with a turbine exhaust shell) is preferably adopted. A well designed exhaust diffuser can improve the efficiency of the turbine stage by about 3-5%. Obviously, under the additional measures that the average flowing direction of the outlet of the nozzle ring vane (or the single-flow-channel or double-flow-channel vaneless volute) of the turbine is matched with the flowing direction of the inlet of the impeller vane (namely the average flowing direction of the outlet of the nozzle ring vane and the impeller vane is approximately vertical to the inlet edge of the vane), and a well-designed exhaust diffuser is arranged at the outlet of the impeller, the mixed flow turbine impeller designed according to the geometric parameter selection range of the mixed flow turbine impeller recommended by the invention has the advantages that compared with the radial flow turbine and the mixed flow turbine in the prior art, the delta R value and the delta value₁The increase of the rotating speed and the adverse effect of the 'secondary flow' on the flow in the wheel are greatly weakened, so that the mixed flow turbine stage designed by the invention is ensured to obtain higher efficiency, namely the effect of the invention.

Drawings

Fig. 1a is a meridian flow line distribution schematic diagram of a flow field in a mixed flow type turbine impeller. The figure shows the position and shape of the meridian flow line (gyration surface bus) forming the hub gyration surface flow surface, the rim gyration surface flow surface, and the middle (average, central) gyration surface flow surface which bisects the in-wheel flow. Wherein, the hub and the rim gyration surface flow surface are respectively formed by the meridian surface profiles of the hub and the rim of the impeller revolving around the axial lead of the rotating shaft of the impeller. The meridian flow line (bus) of the flow surface of the middle gyration surface respectively passes through the geometric mean radius points a and b of the inlet and the outlet of the impeller blade. Fig. 1b is a schematic view of a turbine blade cascade on any gyration surface.

Fig. 2 shows a slope inclination angle δ (arctg/dz) at any point a on any meridian flow line (revolution surface flow surface bus) S in the internal flow field of the mixed-flow turbine impeller. Delta is the included angle between the tangent line at the point A on the meridian flow line S and the axis of the rotating shaft. The radius difference delta r of the meridian flow line at the inlet and outlet of the impeller is (r)₁-r₂)。

FIG. 3 shows the rotation degree of a fluid at any point A when the fluid flows along any surface of revolution

Where the normal of the plane of revolutionDirection component vector and impeller rotation angular velocity vector

Normal line at point A

The relationship of the direction component vectors (the two are turned oppositely, and the magnitude of the rotation component vector is 2 omega sin delta).

FIG. 4 shows that any flow rate is G and the rotation speed is G

Under the working condition, the flow of the blade cascade with any circumgyration surface of the circumgyration can be decomposed into the through-flow of the same blade cascade with one flow G and no rotation (omega is 0) of the fluid circumgyration and the through-flow of the other flow zero (G is 0) but the turning and the rotation

The directions of the partial vectors in the normal direction of the gyration surface are opposite, and the variable strength circulation with the vortex strength of 2 omega sin delta flows and is superposed in the same gyration surface cascade flow passage. The figure shows a schematic of the flow synthesis.

Fig. 5 is a schematic diagram showing velocity superposition of two flows in the turning surface cascade flow channel shown in fig. 4. In the figure 1-flow channel suction surface; and 2, a flow passage pressure surface.

FIG. 6 is a schematic view of a mixed flow turbine stage (consisting of a nozzle vane ring and an impeller) and a meridian section of the impeller with major geometric dimensioning.

Fig. 7a is a schematic structural diagram of a mixed flow turbine impeller suitable for the technology of the present invention. FIG. 7_bIs an example of the meridian section configuration shape of the semi-open impeller of the mixed-flow turbine constructed by the technology of the invention.

Detailed Description

Hereinafter, the technical contents of the present invention will be further described by way of examples with reference to the accompanying drawings.

FIG. 7_aThe structure of the mixed flow type turbine impeller applicable to the technology of the invention is shown schematically. The turbine impeller is a diagonal centripetal turbine impeller, wherein fluid flows into an impeller inlet along a diagonal centripetal manner which forms an angle theta with the radius direction, and then turns to flow out of an impeller outlet axially. Such mixed flow turbine stages have many applications in turbochargers, turboexpanders and small gas turbine plants. FIG. 7_aThe enclosed impeller structure is shown, which is an integral structure formed by combining a wheel cover (wheel rim) 5, blades 4 and a wheel disc (wheel hub) 3 into a whole through precision castingAnd forming an impeller. If the wheel cover 5 is removed, the impeller becomes a semi-open impeller. The closed impeller has high efficiency but low strength; the semi-open impeller has high strength and low efficiency, so the semi-open impeller is widely applied. When the shape of the blade is designed, the inflow angle distribution on the inclined edge of the inlet of the turbine impeller blade must meet the requirement of the impeller inlet speed triangle on the inflow angle distribution on the inclined edge of the inlet under the selected design working condition, so that the large attack angle loss is avoided, and the turbine efficiency is reduced. In general, the relative velocity of the inflow at the geometric mean radius position a of the inlet edge of the vane is selectedAverage meridional flow velocity with the inlet edge of the vane

Are identical, i.e. that

<math> <mrow> <msub> <mover> <mi>W</mi> <mo>&RightArrow;</mo> </mover> <mn>1</mn> </msub> <mo>=</mo> <msub> <mover> <mi>W</mi> <mo>&RightArrow;</mo> </mover> <mrow> <mn>1</mn> <mi>m</mi> </mrow> </msub> <mo>.</mo> </mrow> </math>

Thus, as can be seen from the velocity triangle,

relative inflow angle beta of₁90 ° and the relative speed of the inflow impeller at the hub and rim positionsAnd

relative inflow angle of (2) is due to R_1h＜R₁＜R_1sh(i.e. the

<math> <mrow> <msub> <mover> <mi>U</mi> <mo>&RightArrow;</mo> </mover> <mrow> <mn>1</mn> <mi>h</mi> </mrow> </msub> <mo><</mo> <msub> <mover> <mi>U</mi> <mo>&RightArrow;</mo> </mover> <mn>1</mn> </msub> <mo><</mo> <msub> <mover> <mi>U</mi> <mo>&RightArrow;</mo> </mover> <mrow> <mn>1</mn> <mi>sh</mi> </mrow> </msub> </mrow> </math>

) To respectively make beta_1h< 90 DEG and beta_1shThe distribution requirement of the inlet angle of a forward bending and backward sweeping type is formed on the inclined edge of the inlet of the impeller blade after the angle is more than 90 degrees. The distribution can be realized by additionally implementing beveling of the inlet edge and adjusting the shape of a parabola on a reference surface on a general method for molding radial turbine impeller blades-molding radial ruled paraboloid blades in the prior art. If it is changed again

<math> <mrow> <msub> <mover> <mi>W</mi> <mo>&RightArrow;</mo> </mover> <mn>1</mn> </msub> <mo>=</mo> <msub> <mover> <mi>W</mi> <mo>&RightArrow;</mo> </mover> <mrow> <mn>1</mn> <mi>m</mi> </mrow> </msub> </mrow> </math>

More types of inlet angle distribution variations are obtained by the location of the point on the oblique edge of the inlet of the blade. When the inflow angle of the inclined edge of the blade inlet is adjusted, only the included angle between the tangent line of the central line of the blade profile cut out by a plane vertical to the inclined inlet edge at the inlet edge and the circumferential direction of the impeller is calculated as the relative inflow angle of the position; the influence of the thickness of the blade should not be taken into account, and the included angle between the tangent line of the suction surface (or the pressure surface) of the blade profile at the inlet edge and the circumferential direction of the impeller is taken as the relative inflow angle beta of the position₁. Since the relative inflow angle is calculated tangentially to the centerline of the leading edge of the airfoil, it accounts for both the effect of suction side thickness variations on the profile shape of the airfoil inlet. The influence of the change in thickness on the pressure surface side is also taken into account, so that the influence of the change in thickness on the suction surface side (or pressure surface side) alone on the geometry of the leading edge of the blade profile, or the influence of the change in thickness on the leading edge of the blade profile alone, is taken into accountThe method for calculating the relative inflow angle by the included angle between the tangent of the suction surface (or the pressure surface) and the circumferential direction of the impeller is more reasonable and more practical. Therefore, when the inflow angle adjustment calculation is involved in the impeller blade modeling, only the shape of the ridge surface in the blade needs to be adjusted, and the adjustment of the blade thickness distribution should not be involved.

FIG. 7_bAn example of a radial cross-sectional shape of a semi-open impeller of a mixed-flow turbine constructed in accordance with the present invention is shown, with the following geometric parameters:

for a typical prior art radial turbine wheel design:

D_1sh＝D_1m＝D_1h，D_2sh＝0.85D_1m，D_2h＝0.4D_2sh＝0.34D_1m，

D_2m＝[(D² _2sh+D² _2h)/2]^0.5＝0.64734D_1m，

＝D_2m/D_1m＝0.64734，

<math> <mrow> <mrow> <mo>(</mo> <msubsup> <mi>U</mi> <mn>1</mn> <mn>2</mn> </msubsup> <mo>-</mo> <msubsup> <mi>U</mi> <mn>2</mn> <mn>2</mn> </msubsup> <mo>)</mo> </mrow> <mo>/</mo> <mn>2</mn> <mo>=</mo> <msup> <mi>ω</mi> <mn>2</mn> </msup> <mrow> <mo>(</mo> <mn>1</mn> <mo>-</mo> <msubsup> <mover> <mi>R</mi> <mo>&OverBar;</mo> </mover> <mn>2</mn> <mn>2</mn> </msubsup> <mo>)</mo> </mrow> <msubsup> <mi>D</mi> <mrow> <mn>1</mn> <mi>m</mi> </mrow> <mn>2</mn> </msubsup> <mo>/</mo> <mn>8</mn> <mo>=</mo> <mn>0.0726</mn> <msup> <mi>ω</mi> <mn>2</mn> </msup> <msubsup> <mi>D</mi> <mrow> <mn>1</mn> <mi>m</mi> </mrow> <mn>2</mn> </msubsup> </mrow> </math>

the wheel diameter D of the mixed-flow turbine impeller designed by the invention and the prior art radial-flow turbine impeller_1mUnder the same working condition of the same rotating speed, the centripetal flow in the mixed-flow turbine impeller has much smaller function of overcoming the centrifugal force than that of the radial-flow turbine impeller, the ratio of the centripetal flow to the radial-flow turbine impeller is 0.45: 1, namely the structure of the mixed-flow turbine impeller greatly weakens the adverse effect of the centrifugal force field on the flow in the mixed-flow turbine impeller.

By analyzing this example, it can be seen that

The importance of the value on the influence of the flowfield in the wheel. The impeller structure with large theta value is adopted, although the delta value of the impeller inlet is mainly reduced, the R value is concomitant with the theta value₁Caused by a decrease in value

The beneficial effect of increased values on the in-wheel flow field is likewise not negligible.

From the geometric shape of the flow channel of the meridian section of the impeller in the embodiment, the mixed-flow turbine is an improved form for obtaining high efficiency under the working condition of high specific speed as a radial-flow turbine, and the meridian shape of the impeller (especially the part from the middle gyration surface to the rim gyration surface) of the mixed-flow turbine is very close to that of an axial-flow turbine impeller.

Claims

1. The utility model provides a mixed flow turbine wheel, impeller are closed impeller or semi-open impeller, and closed impeller comprises blade 4, rim plate 3 and 5 triplexes of wheel cap, and semi-open impeller comprises blade 4, rim plate 3, its characterized in that: axial length B of through-flow part of impeller blade and geometric mean diameter D of inlet edge of impeller blade_1m＝2R₁Ratio of

＝B/D_1mIn the range of

0.45-0.60 percent; geometric mean radius R of outlet edge of impeller blade₂＝[(R_2sh ²+R_2h ²)/2]^0.5Geometric mean radius R of inlet edge of impeller blade₁＝[(R_1sh ²+R_1h ²)/2]^0.5Ratio of

＝R₂/R₁In the range of

＝0.73～0.93。

2. The impeller of claim 1, wherein: the selection range of the included angle theta between the inlet edge of the impeller blade and the axis of the rotating shaft is 20-70 degrees; length l of inclined edge at inlet of impeller blade₁Geometric mean diameter D of inlet edge of impeller blade_1mRatio of

＝l₁/D_1mIn the range of

0.16-0.25; the included angle gamma between the outlet edge of the impeller blade and the radius direction is-15 to 30 degrees.