JPH08109801A - Supercharger turbine - Google Patents

Abstract

PURPOSE: To make compact, light in weight a supercharger having a small moment of inertia so as to markedly improve its characteristics by notching a leading edge facing the side of a scroll chamber and forming an intermediate zone in the middle way reaching a turbine housing chamber. CONSTITUTION: A turbine housing chamber 10 for freely rotatably housing a turbine wheel 9 is formed in a turbine casing 2, a scroll chamber 11 for introducing exhaust gas, etc., is partitively forward in its outer periphery, a leading edge part 17 facing the scroll chamber 11 side of the turbine blade 13 of the turbine wheel 9 is notched and an intermediate zone 21 is formed in the middle way from the scroll chamber 11 to the turbine housing chamber 10.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a turbine used in a turbocharger for automobiles and the like.

[0002]

2. Description of the Related Art Conventionally, radial (radial flow) type turbines are predominantly used as turbines for turbochargers for automobiles and the like.

FIG. 14 shows a turbocharger equipped with the radial turbine. As shown, the turbocharger a has a turbine casing b, in which a turbine housing chamber c and a scroll chamber d are formed, and a turbine impeller e is rotatably housed in the turbine housing chamber c. The exhaust gas in the scroll chamber d is introduced into the turbine impeller e to rotate it. The scroll chamber d is formed in the radial direction along the outer periphery of the turbine housing chamber c, so that the exhaust gas in the scroll chamber d flows radially inward and is introduced into the turbine housing chamber c, where the turbine impeller e
Is rotated and the flow direction is changed to a right angle, and thereafter the gas is exhausted to the tip side in the turbine shaft f direction.

[0004]

By the way, in the case of such a radial type turbine, the outer diameter of the turbine impeller e becomes relatively large and the moment of inertia becomes large so that the characteristics as a supercharger, for example, the rising characteristics ( It has the drawback that it worsens the response. Further, the exhaust gas from the scroll chamber d is introduced between the turbine blades g of the turbine impeller e, and particularly, a strong centrifugal force acts on the exhaust gas between the turbine blades g in the gas introduction portion h, This hindered the introduction and caused the deterioration of flow rate characteristics and performance.

On the other hand, in order to prevent the turbine impeller e from increasing in size, a mixed flow turbine has been partially popularized. Specifically, as shown in FIG. 15, the leading edge portion (end edge portion) i of the gas introduction portion h of the turbine blade g is formed obliquely, and the scroll chamber d
Also, an oblique flow path is formed so that the exhaust gas is introduced from the vertical direction of the leading edge portion i, and the exhaust gas is allowed to pass obliquely as a whole to rotationally drive the turbine impeller e. The leading edge portion i of the turbine blade g is formed at a predetermined angle α with respect to the turbine axis j, that is, an angle of 30 ° at the maximum, and the scroll chamber d is adjusted to the angle of the leading edge portion i, Turbine impeller e
It has a shape that squeezes out (overhangs) toward the back side of.

Further, there is a bearing portion (indicated by k in FIG. 14) for the turbine shaft f on the back side of the turbine impeller e, so that if the protrusion of the scroll chamber d is made too large, the bearing portion will be heated by the heat of exhaust gas. Becomes hot and special cooling or heat shielding measures must be taken. Therefore, in the case of the conventional mixed flow turbine, the angle α of the leading edge portion i of the turbine blade g is set to be relatively small,
As a result, the protrusion of the scroll chamber d is reduced.

However, with such a blade shape, the centrifugal force of the exhaust gas at the gas introduction portion h cannot be completely removed, so that the exhaust gas cannot be introduced smoothly. Therefore, the conventional mixed flow turbine must have characteristics close to those of the radial turbine, and as a result, it is impossible to obtain desirable characteristics as a supercharger.

Particularly in these turbines, the scroll chamber d and the turbine accommodating chamber c are in direct communication with each other,
Therefore, the exhaust gas in the scroll chamber d is directly introduced into the turbine housing chamber c and supplied to the rotation of the turbine impeller e.

Therefore, the present invention was devised in view of the above viewpoint, and an object thereof is to make a supercharger compact and lightweight, having a small moment of inertia and capable of greatly improving the characteristics as a supercharger. To provide a turbine for use.

[0010]

In order to achieve the above-mentioned object, a turbine for a supercharger according to the present invention has a turbine housing in which a turbine impeller is rotatably housed in a turbine casing and the outer circumference thereof. A scroll chamber for introducing exhaust gas or the like is formed by partitioning, a leading edge portion facing the scroll chamber side of the turbine blade of the turbine impeller is cut out, and an intermediate zone is provided between the scroll chamber and the turbine storage chamber. It was formed.

[0011]

According to the above, the influence of centrifugal force is eliminated by notching the leading edge portion of the turbine blade.
The intermediate zone also directs the exhaust gas from the scroll chamber towards the leading edge of the turbine blade. Since there is no influence of centrifugal force, the exhaust gas in the intermediate zone can be smoothly introduced.

[0012]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT A preferred embodiment of the present invention will be described in detail below with reference to the accompanying drawings.

FIG. 1 is a side sectional view showing a turbocharger equipped with a turbine according to the present invention. As shown in the figure, the turbocharger 1 has a turbine casing 2, a center casing 3, and a compressor casing 4, which are integrally connected to each other, and a bearing portion 5 is formed in the center casing 3, and The turbine shaft 6 is rotatably supported. Further, the bearing portion 5 is provided on the center casing 3.
A large number of oil holes 7 up to are provided, through which lubricating oil is supplied to the bearing portion 5. A compressor 8 is fixed to one end of the turbine shaft 6 in the compressor casing 4,
A turbine impeller 9 is provided at the other end.
The turbine impeller 9 is housed in a turbine housing chamber 10 formed in the turbine casing 2, and a scroll chamber 11 is formed in the turbine casing 2 in a radial direction with respect to the turbine housing chamber 10. The scroll chamber 11 is formed along the outer periphery of the turbine accommodating chamber 10, and the cross section thereof is sequentially enlarged or reduced along the circumferential direction.

In the turbine impeller 9, a plurality of (8 to 12) turbine blades 13 are erected on the outer peripheral surface of a disk 12 at equal intervals, and the disk 12 has a substantially conical shape and its outer peripheral surface. Is a smooth curved surface. Here, for convenience, the illustrated turbine impeller 9 is represented by a meridional surface shape in which the turbine impeller 9 is cut along a vertical plane including the turbine axis 14 or a vertical plane and the turbine blades 13 are positioned on the plane. The impeller 9, particularly the turbine blade 13, has a more three-dimensional and complicated shape. Usually, the outermost peripheral edge 15 of the turbine blade 13 is called a shroud, the connecting edge 16 of the turbine blade 13 to the disk 12 is called a hub, and the gas introduction side edge 17 of the turbine blade 13 is a leading edge and a gas outlet. The side edge portion 18 is referred to as a trailing edge portion.

2 and 3 are a side view and a partial cross-sectional view showing the shape of the turbine blade 13 more clearly.
First, in FIG. 2, one turbine blade 13a is attached to the shroud 1.
As seen from the view from the 5 side, the turbine blade 13
The gas introduction portion 19 of a is oriented so as to face the rotation direction along the gas flow direction. Then, the turbine blade 13a is smoothly bent back toward the opposite side in the rotational direction at its central portion, and the gas lead-out portion 20 is oriented so as to face the opposite side in the rotational direction. Next, in FIG.
FIG. 3 is a view of the turbine blade 13 as viewed in a vertical cross section of the turbine shaft center 14 along the gas flow direction. In all cross sections A to E, the thickness center line of the turbine blade 13 passes through the turbine shaft center 14. I understand. That is, the turbine blade 13
Is toward the turbine axis 14 when viewed at any cross section along the gas flow direction. This is because of the strength reason, that is, the centrifugal force during rotation that acts on each part of the turbine blade 13 is completely received by the root part of the disk 12. The warped portion of the turbine blade 13 or 13a shown in FIG. 2 is located between the sectional positions b and c in FIG.

In particular, in FIG. 3A, the turbine impeller 9 is represented by a meridional surface shape, and the turbine blade 1
The leading edge portion 17 of No. 3 is notched so as to be inclined at a predetermined angle α with respect to the turbine axis 14 a parallel to the turbine axis 14. In particular, in the case of this embodiment, the angle α (hereinafter referred to as oblique degree) exceeds 35 ° and 60 °.
It is set to a smaller value, and its angle α is set larger than that of the conventional mixed flow turbine (maximum 30 °).

FIG. 4 shows points a to e of FIG. 3 projected on a plane perpendicular to the meridian plane (ie, the paper surface), and thereby the cross-sectional shape of the turbine blade 13 along the gas flow direction is expressed. . In this figure, the angle γ of the normal gas introduction part 19 with respect to the turbine axis 14a is called the rake angle, which will be described later. Is set to. Further, from this figure, it can be seen that the warped vertex of the turbine blade 13 exists between b and c. In addition, in the cross section of the turbine blade 13 shown in the figure, a line along the center of its thickness is generally called a camber line.

The turbine impeller 9 thus formed is mounted on the turbocharger 1, and as shown in FIG. 1, an intermediate zone 21 is formed between the scroll chamber 11 and the turbine accommodating chamber 10. ing. That is, the intermediate zone 21 is
It is a zone from an opening throat portion 22 provided on the turbine housing chamber 10 side and the axial base end side of the scroll chamber 11 to a leading edge portion 17 of the turbine blade 13 of the turbine impeller 9 (between the dashed lines in the figure). Area). And
On the back side of the turbine impeller 9, the turbine casing 2
Heat shield plate 2 sandwiched by the center casing 3 and the center casing 3.
3 is provided, which forms a partition wall 24 on the axially proximal end side of the intermediate zone 21. Particularly in this embodiment, the partition wall 24
5 with reference to FIG.
It is formed so as to form an angle β ₁ that exceeds 40 ° and is less than 85 °. Further, the partition wall 25 on the axially front end side of the intermediate zone 21 is cut off along the radial direction as shown in the drawing, or has a smooth radius. As a result, the intermediate zone 21 is formed in a ring shape over the entire circumference having a generally triangular or trapezoidal cross section.

By summarizing the above, referring to FIGS. 5 and 6, the optimum dimensional ranges of the turbine impeller 9, the intermediate zone 21, and each part of the scroll chamber 11 determined by simulations and experiments are shown. Here, FIG. 5 shows a turbine impeller 9
FIG. 6 is a side sectional view of the turbine blade 13 in a meridional shape, and FIG. 6 shows the camber line 26 at the position of the shroud 15 of the turbine blade 13. In FIG. 5, the partition wall 25 has a smooth rounded shape.

<< Fig. 5 >> α: 35 ° <α <60 ° β ₁ : 40 ° <β ₁ <85 ° d ₂ h: 0.25d ₁ <d ₂ h <0.35d ₁ d ₂ s: 0.8d ₁ < d ₂ s <d ₁ β ₂ : 70 ° <β ₂ <90 ° δ: 10 ° <δ <40 ° l ₁ : 0.3d ₁ <l ₁ <0.5d ₁ l ₂ : 0.3d ₁ <l ₂ <0.5 d ₁ R: 0.02d ₁ <R <0.1d ₁ where α: obliqueness β ₁ : inclination angle of partition wall 24 with respect to turbine axis 14 a d ₁ : maximum outer diameter of turbine impeller 9 d ₂ h: trailing Hub diameter of edge portion d ₂ s: Shroud diameter of trailing edge portion β ₂ : Angle with respect to turbine axis 14 a formed by the partition wall on the axial direction proximal end side near the throat portion 22 of the scroll chamber 11 δ: Similarly, Turbine axis 1 formed by the partition wall on the tip side in the axial direction
4a Angle l ₁ : Length from axially proximal end surface of turbine impeller 9 to shroud axial tip l ₂ : Similarly, length to hub axial tip R: R radius of partition wall 25 << Fig. 6 >> γs: 40 ° <γs <70 ° β ₂ s: 60 ° <β ₂ s <70 ° l ₃ : 0.1d ₁ <l ₃ <0.25d ₁ where γs: rake of leading edge portion 17 Angle β ₂ s: Turning angle l _{3 at} gas outlet 20: Leading edge 17 at warp-back vertex position A
The distances from β ₁ and β ₂ are set based on the inclination of the hub 16 of the gas introduction part 19 and β ₂ > β ₁ to form a smooth surface.

Next, the operation of the turbine of the above embodiment will be described in detail while comparing it with a conventional radial turbine.

FIG. 7 is a view of the turbine impeller 9r of the radial turbine as viewed from the tip end side in the axial direction, and FIG. 8 is shown in FIG.
The cross section of the turbine blade 13r along the same gas flow direction is shown. In FIG. 8A, the turbine impeller 9r is represented by a meridional surface shape.

As shown in the figure, the leading edge portion 17r of the turbine blade 13r coincides with the direction of the turbine axis 14r, that is, the inclination α is zero. Further, for the above-mentioned reason, the turbine blade 13r faces the turbine axis 14r at any cross section along the gas flow direction. Then, the turbine blade 13r is smoothly bent along the gas flow direction to the opposite side in the rotational direction, and the gas discharge portion 20r is oriented so as to face the opposite side in the rotational direction.

In this case, referring to FIG. 7 and FIG. 9, the exhaust gas has an absolute velocity c before the leading edge portion 17r, while the leading edge portion 17r has a peripheral velocity u.
The exhaust gas is substantially introduced into the gas introducing portion 19r of the turbine blade 13r at a relative speed w.
However, in the case of the radial turbine, the blades 13r cannot be directed in the direction of the relative speed w of the exhaust gas,
Therefore, it was a collision inflow and caused a loss. Further, as described above, a strong centrifugal force is applied to the gas between the blades 13r in the gas introduction portion 19r, that is, the inflow velocity is reduced and the pressure is increased, which hinders the introduction of new gas.

On the other hand, in such a turbine, as shown in FIG. 4, the gas introduction portion 19 of the turbine blade 13 is oriented in the direction of the relative velocity w of the exhaust gas, and thus a collision-free inflow is achieved. Inflow or introduction can be performed smoothly. Further, in particular, the leading edge portion 17 is largely cut out obliquely, and the degree of inclination α thereof is set to a relatively large value of 35 ° <α <60 °. In this way, the inflow speed can be increased and the pressure can be reduced.

Further, the leading edge portion 17
Since the intermediate zone 21 is formed on the front side, this region can be used as a buffer or a transition region to smoothly guide the gas. That is, the intermediate zone 21 obliquely directs the gas of relatively high pressure and high speed from the scroll chamber 11 located in the radial direction toward the leading edge portion 17 of the turbine blade 13 (changes the flow direction), and further The blade 13 is smoothly guided to the gas introduction portion 19. The intermediate zone 21 enables the scroll chamber 11 to be positioned in the radial direction, and prevents the scroll chamber 11 from protruding toward the bearing portion 5 side.

As described above, since the introduction of the exhaust gas is smooth, the high exhaust gas pressure from the engine can be efficiently converted into the rotational force, which can improve the flow rate and efficiency characteristics of the supercharger. And as a result, FIG.
As shown in FIG. 6, the equivalent product 9 can be made smaller than the turbine impeller 9r of the conventional radial turbine. Therefore, the weight and the moment of inertia can be reduced, and the supercharger characteristics such as the rising characteristics can be improved. As a result of making a prototype of such a turbine impeller 9, the moment of inertia was able to be about 50% compared with the conventional product.

FIG. 11 shows a conventional radial type turbine and
In the graph showing the comparison example with the turbine,
Is the pressure ratio, that is, the ratio of the inlet pressure to the outlet pressure, and the horizontal axis
It has a flow rate. Also here, turbine impeller
Rotation speed n of₁, N₂, N ₃And each constant, and
I use it as a parameter.

First, in the radial type shown in (a), due to the influence of the centrifugal force of the inflowing gas, the relationship between the pressure ratio and the flow rate (broken line) changes stepwise with respect to the change of the rotation speed n, and the rotation speed n.
Even if the value rises, the flow rate corresponding to it cannot be flowed.

On the other hand, in the case of such a turbine shown in (b), the relationship (solid line) changes continuously according to the change of the rotation speed n, that is, the flow rate continuously increases as the rotation speed n increases.

As described above, in the case of such a turbine, the relationship between the pressure ratio and the flow rate is unlikely to be influenced by the rotational speed, and as a result, a large flow rate can be made to flow in a region where the pressure ratio is high, which is very suitable as a supercharger. It is possible to obtain good flow characteristics.

Further, the maximum turbine efficiency obtained by the conventional radial type was about 70%, but in the case of such a turbine, it was 77 to 79%, showing a dramatic improvement.

FIG. 12 shows the relationship between the turbine inlet pressure and the engine crank angle. As shown, the inlet pressure varies with the crank angle. The broken line is the average pressure. According to this, the pressure becomes maximum at the point A, but the dimensions of each part of the turbine are determined so that the collision-free inflow can be mainly achieved at the pressure at the point A. Therefore, the maximum exhaust gas pressure of the engine can be effectively used, and the turbine can have a very high performance.

FIG. 13 shows that α, γs,
It is a matrix table (star chart) showing the performance test results summarized for β ₁ . In this case, α = 0 °, β ₁
= 90 ° is completely radial and is excluded.
Further, ◯ means good, Δ means performance close to radial type, and x means equal to radial type. The range of each value mentioned above is 35
° <α <60 °, 40 ° <γs <70 °, 40 ° <β ₁ <85 °, and it can be seen from the table that they are suitable. still,
In the case of α = 60 ° and β ₁ = 40 °, 50 °, 60 °, the protrusion of the scroll chamber 11 becomes large, which is structurally impossible.

As described above, such a turbocharger turbine exhibits excellent performance or characteristics, and by mounting this turbine, it is possible to dramatically improve the performance of the supercharger or the engine. Although this turbine was designed for a turbocharger of an automobile engine, it is of course possible to apply such a configuration to a gas turbine, an expansion turbine or the like.

On the other hand, the conventional radial turbine has another problem as described below.

As shown in FIG. 18, as a turbine impeller 30 of a radial turbine, a disk 31 thereof is generally shaped like a chain double-dashed line, that is, the disk back surface 32 is a turbine blade. The disk surface 33 has a size such that 33 is hidden, and the hub surface 33a to which the turbine blade 33 is attached extends to the maximum radial position of the turbine blade 33. However, in this case, particularly when rapid heating / cooling associated with acceleration / deceleration is frequently repeated, such as for automobiles, a relatively thin turbine blade 33 with a small heat capacity and a disk 31 with a relatively large heat capacity are used. A large thermal deformation difference occurs between the turbine blades 33 and the turbine blades 33 are destroyed. That is, at the time of rapid heating, the turbine blade 33 tries to expand immediately but the disk 31 does not extend, and at the time of rapid cooling, the turbine blade 33 tries to shrink immediately but the disk 31 does not shrink.
As a result, large thermal stress is generated at the root portion of the turbine blade 33, which is located radially outward, and destruction occurs. Here, in general, the amount of thermal deformation
The thermal stress is large, and conversely the amount of thermal deformation
It is known that thermal stress is small and no fracture occurs.

Conventionally, as a solution to this problem, as shown by the solid line in FIG. 18, a scallop 34 is provided by notching the radially outer portion of the disk 31 located between the turbine blades 33. The scallop 34 is a groove-shaped space formed between the turbine blades 33, and this reduces the radius of the disk back surface portion 32. When viewed from the back surface portion 32 side, the turbine blade 3 as shown in FIG.
It looks like the radially outer part of 3 is protruding. When the scallop 34 is provided, the turbine blade 33
A thick portion 35 is formed at the base of the protruding portion so as to leave the shape of the two-dot chain line in FIG.
The rear end surface of the projecting portion including is flush with the disk rear portion 32 and extends in the radial direction. Thick part 35
Has a skirt shape as shown in (a), and has a shape like a web. When the scallop 34 is provided in this manner, the radially outer portion of the disk 31 can be deleted to allow the turbine blade 33 to freely expand and contract, and prevent destruction. In addition, during expansion and contraction due to rapid heating and rapid cooling, the thick portion 35 serves as a buffer portion to relieve thermal stress.

However, the provision of such a scallop 34 causes the following problems.

As shown in FIG. 19, the exhaust gas introduced from the scroll chamber 36 has a gap 38 between the exhaust gas P flowing in from the regular inlet toward the inner side in the radial direction and the turbine casing 37 from the back side of the turbine blade 33. And Q coming in from the scallop 34 through the scallop 34. In particular, the latter Q only blocks the inflow of the normally inflowing P and does not contribute to the rotation of the turbine impeller 30. Further, since the protruding portion of the turbine blade 33 stirs the gas around the gap 38, this becomes a resistance and the driving energy of the turbine is lost or consumed. In this way, the scallop 34 causes the turbine efficiency to decrease. further,
In order to reduce the inflow gas Q from the scallop 34 as much as possible, it is necessary to pay particular attention to the shape of the gap 38 or the turbine casing 37 (in the illustrated example, it is formed in a crank shape). The variation was directly related to the variation in performance, and the management of the variation was difficult. In particular, this was remarkable in the small turbine in which the gap 38 was inevitably smaller than that in the relatively large turbine.

Therefore, the turbine according to the present invention solves the above problems all at once and greatly contributes to the improvement of turbine efficiency.

As shown in FIGS. 16 and 17, the leading edge portion 43 of the turbine blade 42 of the turbine impeller 41 is inclined with respect to the turbine axis 44 at an inclination α, as described above, and particularly the leading edge portion 43. Is formed to have the same thickness as the other portions, that is, the thickness of the turbine blade 42 which is relatively thin. In other words, the above-mentioned thick portion is not provided. The root end 46 of the root portion 45 is connected to the outermost peripheral edge 50, which is the position where the hub surface 48 of the disk 47 and the back surface portion 49 intersect. The partition wall 52 on the axially proximal end side of the intermediate zone 51 is shaped according to the shape of the hub surface 48 of the disk 47 or according to the line thereof. In particular, in FIG. 17, the leading edge portion 43 is inclined with respect to the radial direction because the rake angle described above is provided.

According to this structure, the scallop is completely eliminated, and the exhaust gas flowing from the scroll chamber 53 through the intermediate zone 51 is entirely introduced from the leading edge portion 43 which is a regular inlet. Further, since the gap 55 formed between the turbine casing 54 and the turbine casing 54 is the gap between the rear face portion 49 of the disk 47, the gas passing therethrough disappears, and the flow of the gas that normally flows in is not blocked.
Further, since the disk-shaped disk back surface portion 49 rotates facing the gap 55, the gas in the gap 55 is not agitated and no resistance is generated. In particular, the resistance difference between the scratching with the disk-shaped disk back surface portion 49 and the scratching with the protruding portion of the turbine blade as in the conventional case is significantly different. With these, it is not necessary to strictly set the shape of the turbine casing 54 and strictly control the dimensions of the gap 55.

As described above, the cause of the decrease in turbine efficiency due to the scallop is eliminated, and the energy of the exhaust gas can be used extremely efficiently, and the turbine efficiency can be improved.

Further, since the root portion 45 of the leading edge portion 43 is located on the inner side in the radial direction, there is no fear of thermal stress destruction due to rapid heating / cooling even when it is intended for an automobile, and the turbine blade 42 freely expands and contracts. It is possible and the thick part is not required.

Further, in particular, the inclination α is set to a relatively large angle exceeding 35 ° as described above, and the diameter of the disk back surface portion 49 (the diameter formed by the outermost peripheral edge position 50) d ₁ h and the maximum of the turbine impeller 41 are set. The relationship with the outer diameter d ₁ is d ₁ h ≦ 5/6 × d ₁
Then, the root portion 45 or the root end 46 can be positioned sufficiently inward in the radial direction, and the need for the scallop and the thick portion can be completely eliminated. It has also been confirmed by experiments that it is preferable to set α to 40 ° to 45 °. Conventionally, in mixed-flow turbines that are mainly used for marine turbochargers, the maximum degree of inclination is 30 °, so a scallop is required when using this for automobiles and motor boats. . Therefore, the turbine according to the present invention is superior to the conventional mixed flow turbine in terms of prevention of destruction, improvement of efficiency, and the like.

[0047]

The present invention exhibits the following excellent effects.

(1) The gas can be smoothly introduced, and the characteristics of the supercharger can be dramatically improved.

(2) The size can be reduced, and the weight and the moment of inertia can be reduced.

(3) In particular, according to the present invention as set forth in claim 6, it is possible to improve the turbine efficiency by eliminating the inflowing gas from the scallop.

[Brief description of drawings]

FIG. 1 is a side sectional view showing a turbocharger including a turbine according to the present invention.

FIG. 2 is a side view showing a turbine impeller.

FIG. 3 shows a turbine impeller, (a) shows its meridional surface shape, and (b) shows a cross section of the turbine blade along the gas flow direction.

FIG. 4 is a view showing a cross-sectional shape of turbine blades along a gas flow direction.

FIG. 5 Turbine impeller (meridian shape), intermediate zone,
It is a side sectional view showing a scroll chamber.

FIG. 6 is a diagram showing a camber line at a shroud position of a turbine blade.

FIG. 7 is a view showing a turbine impeller of a conventional radial type turbine, particularly when viewed from the tip side in the axial direction.

FIG. 8 shows a turbine impeller of a radial turbine,
(A) shows the meridional surface shape, and (b) shows the cross section of the turbine blade along the gas flow direction.

FIG. 9 is a view showing a cross-sectional shape of turbine blades along a gas flow direction in a radial turbine.

FIG. 10 is a diagram showing a comparison of turbine impellers.

FIG. 11 is a graph showing the relationship between pressure ratio and flow rate,
(A) is the case of a radial type turbine, (b) is the case of such a turbine.

FIG. 12 is a graph showing how the turbine inlet pressure varies with the crank angle of the engine.

FIG. 13 is a matrix table (star chart) showing performance test results.

FIG. 14 is a view showing a conventional example, and is a side sectional view showing a turbocharger including a radial turbine.

FIG. 15 is a view showing a conventional example and is a side sectional view showing a mixed flow turbine.

FIG. 16 is a side sectional view showing a turbine according to the present invention.

FIG. 17 is a schematic view of the turbine impeller of FIG. 16 when viewed from the back side.

FIG. 18 shows a turbine impeller of a conventional radial turbine, (a) is a schematic view seen from the rear side, (b).
FIG.

FIG. 19 is a side sectional view showing a conventional radial turbine.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 Turbocharger 2 Turbine casing 9 Turbine impeller 10 Turbine accommodation chamber 11 Scroll chamber 13,42 Turbine blade 14a Turbine axis 15 Shroud 17,43 Leading edge part 19 Gas introduction part 21 Intermediate zone 24 Partition wall 46 Base root 47 Disk 50 Outermost edge γs Rake angle of leading edge part α Angle formed by leading edge part with respect to turbine axis (oblique degree) β ₁ Inclination angle of partition wall with respect to turbine axis

Claims (6)

[Claims] 1. A turbine housing chamber for rotatably housing a turbine impeller is formed in a turbine casing, and a scroll chamber for introducing exhaust gas and the like is defined on the outer periphery of the turbine housing chamber. Cut out the leading edge facing the scroll chamber side,
A turbocharger turbine, characterized in that an intermediate zone is formed between the scroll chamber and the turbine accommodating chamber. 2. The turbocharger turbine according to claim 1, wherein the scroll chamber is formed in a radial direction with respect to the turbine housing chamber. 3. The leading edge portion of the turbine blade is cut out so as to be inclined so as to form an angle of more than 35 ° and less than 60 ° with respect to the turbine axis in a meridional surface shape. A turbocharger turbine according to claim 1. 4. The turbocharger turbine according to claim 1, wherein a rake angle of the leading edge portion of the turbine blade exceeds 40 ° and is smaller than 70 °. 5. The partition wall located on the turbine axial direction base end side of the intermediate zone is 40 ° with respect to the turbine axis line.
The turbocharger turbine according to any one of claims 1 to 4, wherein the turbocharger turbine is formed so as to be inclined at an angle exceeding 85 ° and less than 85 °. 6. The turbocharger turbine according to claim 1, wherein a root end of the leading edge portion of the turbine blade is connected to an outermost peripheral edge of a disk.