Exhaust-gas Driven Turbochargers

Description

SUMMARY OF THE INVENTION

The invention concerns a waste or exhaust gas turbo-charger for use in connection with internal combustion engines in which a turbo-charger is operatively connected a turbine- and a compressor wheel, preferably via a common drive shaft.

Known waste gas turbo-superchargers for internal combustion engines are adjusted to the average energy per time unit of the exhaust gases which emanate from respective cylinders of an internal combustion engine.

With relatively low outputs and correspondingly small amounts of exhaust gases, the turbine of the gas turbo-charger is designed as a "centripetal turbine," which is traversed in radial direction from the outside or exterior to the inner part of the turbine blade wheel from where these gases issue axially. These centripetal turbines, however, are not adapted to a periodically greatly varying or changing energy supply in order to achieve optimal employment thereof. The form of the blade wheel, in particular the form of its blades and those of the compressor wheel, is designed and developed according to certain viewpoints taking into consideration only turbines to which gases of constant energy are fed and administered.

Compressor wheels of known gas turbo-superchargers of this kind are designed generally as radial wheels whose diameters according to the state of the art should be as far as possible equal to the diameter of the turbine wheel.

It is therefore an object of the invention to provide means overcoming heretofore known constructional disadvantages and considerably improving the design of gas turbo-superchargers in such a way that varying energy supply is reckoned with, and its utilization is considered for the purpose of greater efficiency during the operation of these engine cylinders.

Besides these considerations and advantages resulting therefrom, the gas turbo-supercharger according to the invention is lighter in weight and smaller than known structures of gas turbo-superchargers. The reduction in weight has the further advantage that smaller quantities of expensive construction material is employed, so that the exhaust gas turbo-supercharger according to the invention is more economical and less expensive to produce than known exhaust gas aggregates of this type.

Gas turbo-superchargers for supercharging piston-driven internal combustion engines, where the turbine- and the compressor wheels are arranged on a common shaft, are known in the art, but there remains still the problem to be solved pursuant to this invention, namely, that the turbine wheel is to be traversed diagonally from the exterior to the inner part thereof so that the inlet edges of the blades of the turbine wheel, projected in meridian planes, extend substantially perpendicularly to the direction of the entering gas current, projected into the same meridian plane, and further, that the blade cross sections of the turbine wheel are radially arranged in sectional planes normal to the axis. The blades of the turbine wheel form at the inlet edge an entrance angle of less than 90.degree., while the turbine housing is designed as a single-flow spiral housing or casing admitting the contemplated medium or fluid to the turbine wheel over its entire circumferential surface, on whose entrance part are joined at least two exhaust gas pipe lines emanating from the cylinders of the internal combustion engine in such a way that the inlet cross-section at a tongue-shaped guide edge of the single flow spiral casing is only slightly larger than the cross-section at the discharge end of each individual exhaust gas pipe lines.

These and other objects and advantages result from the provision of means affording a highly economical exhaust gas and like turbo-supercharger, which is reduced in dimensions and does not take up any considerable space.

The various features of novelty which characterize the invention are pointed out with particularity in the claims annexed to and forming a part of this disclosure. For a better understanding of the invention, its operating advantages and specific objects attained by its uses, reference should be had to the accompanying drawing and descriptive matter in which there is illustrated a preferred embodiment of the invention.

BRIEF DESCRIPTION OF THE ATTACHED DRAWINGS

In the drawings:

FIG. 1 is an axial section through an exhaust gas turbo-charger embodying the invention.

FIG. 2 shows, in a diagram, pressure conditions of the exhaust gas arriving at the turbine as related to the crank angle.

FIG. 3 is a schematic representation of the entrance part of a set of blades of a known radial flow turbine and its associated velocity vector diagram.

FIG. 4 illustrates schematically the dependence of the efficiency on the characteristic of a radial flow-turbine according to FIG. 3.

FIG. 5 is a schematic representation of the entrance part of a blade set for a diagonal-flow turbine pursuant to the invention, as well as the velocity vector diagrams associated therewith.

FIG. 6 shows, in a diagram, the dependence of the efficiency on a speed coefficient of an exhaust gas turbo-charger according to FIG. 5.

FIG. 7 is a section taken along the line 7 -- 7 of FIG. 1;

FIG. 8 is a section taken along line 8--8 of FIG. 7.

FIG. 9 shows a detail on an enlarged scale of a left-hand portion of FIG. 1, to which reference is had in the description.

FIG. 10 illustrates a velocity vector diagram derived from FIG. 9.

FIG. 11 shows, in development, a portion of the blade wheel of FIG. 9.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS

Referring now more specifically to the drawings and, in particular, to FIG. 1, there is disclosed a turbine wheel 1 with blades 2. The broken line 3 denotes the path of a central filamentous flow line, (circularly projected in axial section of FIG. 1). Gas admission to turbine wheel 1 occurs diagonally in the direction 36. The flow passage likewise takes place substantially diagonally to the represented meridian section.

The turbine wheel 1 projects, at the inlet edge 4 of blade 2, by an amount LU in the direction of bearing 14 with respect to the rear wall 5 of wheel 1, so that the wheel length L is extremely small in the direction of the axis and the center of gravity of wheel 2 is displaced toward bearing 14 of shaft 15. This arrangement causes sufficient quietness in the rotor run and operation. The turbine wheel is thus lighter in weight so that it is less expensive, because of the high price of its employed high temperature-resistant material.

The inlet edge 4 of the wheel is inclined by the angle .beta..sub.3 relative to the direction of the axis of rotation. This has the effect that:

1. the mean diameter D1 at the inlet edge 4 is kept smaller than the maximum diameter D.sub.2 at the outlet edge 16; and further

2. the length L of the wheel is reduced by the difference between the length of the inlet edge 4 and the measure of projection in respect to the axis thereof, so that material will be saved.

If a lot of energy is available at the entrance into the turbine, that is, if the pressure gradient in the turbine is great and if large amounts of exhaust gas pass or flow through the turbine, the velocity, with which the gas flows out of blades 2, is very high. Such high outlet velocities must be allowed in order to obtain very small wheel diameters and thus less expensive turbine wheels. The circumferential component of this high outlet velocity can be utilized in a radial diffuser 6. By recovering a portion of the outlet velocity, a pressure below atmospheric pressure is obtained in the space between outlet edge 16 and closure cover 9.

The outside diameter D2 of outlet edge 16 of turbine wheel 1 is the maximum diameter of wheel 1. This has the advantage that the spiral casing 7, if necessary together with the discharge spiral casing defining a diffuser 6, can be connected rigidly or difficulty detachably with bearing housing 8 or even be made integral with the latter. Nevertheless, the turbine wheel 1 can be disassembled in the direction toward the right side. Cover 9 may be kept relatively small. Under certain circumstances, wheel 1 can also be taken off through a centrally arranged short exhaust pipe flange (not shown).

In the diagonally traversed turbine wheel 1, according to the invention, the mean diameter D1 is smaller than the outside diameter (maximum diameter) D2 of outlet edge 16. This is due to the fact (as pointed out herein above) that the inlet edge 4 is obliquely cut off at an angle .beta..sub.3 . This has the advantage that, at an equal cross sectional area 25 of spiral casing 7,the radius Rs of the center of gravity of this cross-sectional area 25 diminishes. The spiral casing 7 is thus kept smaller and cheaper in its construction.

The left portion of FIG. 1 shows the compressor wheel 37 with blades 12 for air flow traversal from the inside toward the peripheral edge 13 or outside. The mean diameter D12 of outlet edge 13 of blades 12 should not be greater than 88 percent of the outer diameter (maximum diameter) D2 of the outlet edge 16 of turbine wheel 1.

The compressor wheel 37, as shown, is traversed by air diagonally, i.e., it is administered axially, the flow taking place between edge 26 and outlet edge 13 in diagonal direction 38. The outlet edge 13 is so inclined relative to the direction of the axis of rotation by an angle .beta.4 that it extends perpendicularly to the direction 38 of the escaping or outflowing air. In the same manner as indicated for the turbine wheel 1, in the right hand part of FIG. 1, the outlet edge 13 of blade 12 of compressor wheel 13 could also project, with respect to the rear wall of this wheel, by a certain amount in the direction of bearing 14 so as to obtain an extremely small axial length of the wheel, a displacement of the center of gravity of blades 12 toward the bearing 14 of shaft 15 and an improved quietness during running of compressor wheel 37.

According to FIG. 1, the maximum diameter (outer diameter) D11 of inlet edge 26 of blades 12 of compressor wheel 37 is also greater than the mean diameter D12 of outlet edge 13. This has the advantage as already pointed out, that the maximum diameter (outer diameter) D2 of the outlet edge 16 of the turbine wheel is greater than the mean diameter D1 of the inlet edge 4.

FIG. 2 shows the pressure P in front of the turbine as a function of the crank angle (from 630.degree. to 180.degree.) of one of the cylinders of the internal combustion engine connected to the exhaust gas turbo-supercharger. Each individual cylinder yields a pressure course with periodically recurring maximum and minimum. The entire pressure course ahead of the turbine is obtained by superpositions of these pressure curves present in the individual cylinders. The course of the curve can be approximated by a step-like curve whose steps are designated with A, B, C, and D. The upper dead center is designated with OT and the lower dead center with UT.

Based on the periodically changing energy supply, according to FIG. 2, the utilization of the energy supplies for a known blade cascade (FIGS. 3, 4) and for a blade cascade according to the invention (FIGS. 5, 6) are compared.

FIG. 3 shows schematically, in the lower portion, the entrance portion of the cascade of a conventional radial flow turbine (centripetal turbine). In the upper part of FIG. 3, there is shown the velocity vector diagram at the inlet edges of the blades for two different pressures prevailing forward of the turbine, namely, for the values A and C. C is,according to FIG. 2, the mean pressure. The known radial flow turbine is constructed according to this mean pressure C.

With this mean differential pressure C, one obtains a velocity of gas flow C1C, taking into account the circumferential velocity U1 of the inlet edges of the blades, so that a relative velocity W1C is obtained, which has the direction of the entrance portion of the blades (that is, an entrance angle .beta.1 of 90.degree.). The entrance of the gas current tangential to the entrance portion of the blades, which is associated with the mean pressure C, has the result that there is no entrance surge, so that an orderly flow can take place between the blades. Accordingly, the mean pressure C at the maximum efficiency lies on the associated efficiency curve of FIG. 4, which indicates the efficiency K relative to the coefficient U1/co, where U1 denotes the circumferential velocity of the turbine wheel at the outer diameter thereof, and co the theoretic absolute approach velocity of the wheel corresponding to the pressure difference.

The maximum pressure A, however, lies on this curve at a far lower efficiency. In the upper portion of FIG. 3, the absolute velocity of the gas for this maximum pressure A is designated with C1A. Taking into account the unchanged circumferential velocity U1, one obtains a relative velocity W1A. Since this relative velocity W1A forms an angle with the tangent of the blade, an entrance surge appears so that no orderly flow can take place between the turbine blades. This is a main reason why the maximum energy stage A can only be utilized with a low efficiency in known turbines. From FIG. 4 a further reason for a low efficiency K of known exhaust gas turbines may be gathered: the efficiency K of exhaust gas turbines depends on the coefficient U1/co. Because the highest pressure stage A is at a relatively low coefficient U1/co, the total efficiency of known waste or exhaust gas turbines is relatively low.

The aim of the present invention is therefore to improve the energy yield of the turbo-supercharger.

Since the entrance angle .beta.1, according to FIG. 5, is less than 90.degree. according to the invention (preferably 45.degree. to 70.degree.), the turbine wheel blades 2 are so inclined at their entrance portions 2a that this angle of inclination is equal to the angle .beta.1 of the relative velocity W1A at the time of arrival of the energy surge (corresponding to the highest pressure stage A). Furthermore, the circumferential velocity U1 at the inlet edges of the blades is,according to the invention, higher than the corresponding circumferential velocity U1 of a known radial wheel according to FIG. 3.

This has the result, according to FIG. 6, which shows, corresponding to FIG. 4, the output or efficiency K as a function of the coefficient U1/co, that the highest pressure stage A is within the range of the maximum efficiency.

Furthermore,according to the invention the entrance edge 2a of each blade 2 is given a profile which is streamlined. As a consequence thereof, the degree of efficiency K is still very high even at small deviations of the entrance angle from the ideal entrance angle .beta.1 in respect to the relative velocity W. In the efficiency curve according to FIG. 6, this can be seen to show that the range of the maximum degree of efficiency of the impeller wheel according to the invention is much wider than that of the maximum degree of efficiency K of a known execution of a radial wheel according to FIG. 3.

According to the invention, the blades 2 are curved in the range of their inlet edges 2a in cylinder sections and so cut off (angle .beta.3) that the inlet edge 4 in the meridian plane is substantially perpendicular to the relative velocity W1 (direction 36). Consequently, the blade cross sections can be arranged radially in sectional planes perpendicular to the axis, though the entering angle .beta.1 is less than 90.degree..

In radially approached blade wheels (of the above-mentioned centripetal turbines), the blade cross sections cannot be arranged radially in sectional planes perpendicular to the axis with entering angles under 90.degree.; this radial arrangement has advantages with regard to the strength of the wheel which can otherwise be obtained only with a radial wheel having blade entering edges at right angles.

FIG. 7 shows a section taken along line 7--7 through the right half of the turbine of FIG. 1. In FIG. 7 there is only represented the mean radius r1 of wheel 1 which is equal to half the diameter D1 shown in FIG. 1.

FIG. 8 shows a section tbrough FIG. 7 along line 8--8.

In the embodiment of FIGS. 7 and 8, it is assumed that two exhaust gas lines 40 and 41 come forth from the engine, which are to be so combined at the turbine entrance housing, designed as a spiral casing 7, that there exists a very favorable utilization of the intermittently pulsating energy, namely:

a. a combination of the exhaust flows generally just ahead of the turbine in order to conduct the impact in full strength to the turbine, and

b. an admission over the entire turbine circumference to obtain the best efficiency.

To this end, the turbine entrance housing is designed as a single-flow spiral casing 7. FIG. 8 shows how the cross sections Fp of the two gas lines 40 and 41, which are separated by the partition 40a, pass over into the feed cross section Fsp of the single-flow spiral casing 7 on tongue-like guide edge 30. The feed cross section Fsp at tongue edge 30 of the single-flow spiral casing 7 should be equal to each of the cross sections Fp of the two exhaust gas conduits or lines 40 and 41, so that the velocity Cp in cross section Fp passes over unchanged, that is, without any substantial delay or acceleration and without any substantial change in direction, directly into the velocity c1 (see also FIG. 5) in the feed cross section Fsp of spiral casing 7. In practice, it was found feasible for constructional reasons to make Fsp by 5 to 15 percent larger than Fp; these values, however, should not be exceeded. The velocity Cp of the exhaust gas inside the exhaust gas line 40 or 41 must be so high that no exhaust gas can flow back from the one gas line just carrying a gas pressure impact into the other gas line.

The smallest distance r.sub.o of the direction of the gas current entering with the velocity Cp from the axis of rotation should be at most equal to the 1.0 - to the 1.4 fold of the mean wheel radius r.sub.1 in order to avoid greater velocity variations of the waste or exhaust gas current until it enters wheel 1.

FIG. 9 shows a detail of the left-hand portion of FIG. 1, hence from the compressor wheel. This compressor wheel 37 carries wheel blades 12 with the beveled or obliquely cut off (.beta.4) outlet edge 13. FIG. 10 shows a velocity vector diagram at point E of outlet edge 14. U.sub.2 is the circumferential velocity at point E, C2 the absolute air outlet velocity, and W2 the relative velocity of the outflowing air. One realizes that the exit angle .beta.2 is, according to the invention, less than 90.degree., i.e., the blades are inclined toward the rear at outlet edge 13 and against the circumferential direction. Compressor blades inclined toward the rear at the outlet (.beta.2 less than 90.degree.) have the advantage that the charging pressure depends less on the volume current than in blades with radially terminating ends.

For reasons of strength, the blades of the compressor wheel 37 are so designed that the blade cross sections are arranged radially in sectional planes perpendicular to the axis. That this is possible in the compressor wheel traversed diagonally according to the invention, in contrast to the conventional radial wheels, though .beta.2 is less than 90.degree., will be explained on the basis of FIGS. 9, 10 and 11, where FIG. 11 shows a portion of the circumference of the wheel of FIG. 9 in developed form.

E is the point of outlet edge 13 on blade root 42. In the meridian section seen in FIG. 9, the component W2m of the relative exit velocity of the air appears at this point E. The velocity vector diagram represented in FIG. 10 was turned about W2m into the plane in which the velocity W2, with which the air issues from the compressor wheel (relative velocity), appears in correct size. The absolute exit velocity C2 and the circumferential velocity U2 at point E, as well as the angles, appear also in FIG. 10 in their correct size. It is assumed that blade root 42 is straight, in the represented meridian section, between the points E' and E. From the projection of E' into the plane of the velocity vector diagram (FIG. 10) results in FIG. 10 the direction of W2, that is, the angle B2, which is less than 90.degree.. Nevertheless, point F, for example, is arranged radially above E', as it is illustrated in FIG. 11, that is, the blade cross sections are arranged radially in sectional planes perpendicular to the axis.

While a specific embodiment of the invention has been shown and described in detail to illustrate the application of the principles of the invention, it will be understood that the invention may be embodied otherwise without departing from such principles.

Claims

What we claim is:

1. An exhaust gas turbocharger, for use with internal combustion engines, comprising turbine and compressor wheels arranged on a common shaft; said turbine wheel having guide blades structured to be traversed by gases diagonally from the outside of said turbine wheel to the inside thereof; said blades having inlet edges substantially inclined, so that the direction of an entering gas flow, as projected in a meridian plane, is perpendicular to said inlet edges, as projected in the same meridian plane; the cross-section of said turbine wheel guide blades being arranged radially in sectional planes perpendicular to the turbine wheel axis; said turbine wheel guide blades forming, at the inlet edge, an entrance angle of less than 90.degree. with respect to the turbine wheel axis; a housing for said turbine wheel in the form of a single-flow spiral casing for gas admission to said turbine wheel over the entire circumferential surface thereof; and at least two gas conduits connecting respective engine cylinders into said single-flow spiral casing, said gas conduits being joined at said spiral casing in a manner such that the combined inlet cross section of said conduits into said apiral casing is only slightly larger than the cross section of each gas conduit at the ends thereof connected to said spiral casing.

2. An exhaust gas turbocharger according to claim 1, in which the blades (2) of the turbine wheel have at the inlet edge (4) an entrance angle between 45.degree. and 70.degree..

3. An exhaust gas turbocharger, according to claim 1, which in that the inlet edges (4) of said blades (2) of the turbine wheel assume an aerodynamic form and are streamlined in profile.

4. An exhaust gas turbocharger, according to claim 1, which in that the outside diameter (D2) at the outlet edge (16) of each blade (2) of said turbine wheel (1) is the maximum diameter of the turbine wheel.

5. An exhaust gas turbocharger according to claim 1, which in that the inlet cross section (Fsp) at the inlet (3) of said single-flow spiral casing (25) is by 5 to 15 percent larger than the cross section (Fp) at the end of each of said gas lines coming from said engine.

6. An exhaust gas turbocharger, according to claim 1, which in that the smallest distance ro of the direction of the entering gas flow from the axis of rotation is at most equal to 1.0 to 1.4 times the mean wheel radius (rl) at the inlet ledge (4) of said blades (2) of said turbine wheel (1)

7. An exhaust gas turbocharger according to claim 1, in which the distances (Rs) of the centers of gravity of the cross sections of the volute casing (7) from the shaft axis of rotation are kept as small as possible in all meridian sections.

8. An exhaust turbocharger according to claim 1, in which a radial diffuser (6) communicates with the wheel outlet of the turbine.

9. An exhaust gas turbocharger according to claim 1, in which the spiral turbine casing (7) and radial diffuser (6) are integrally connected with a bearing housing (8).

10. An exhaust gas turbocharger according to claim 1, in which the mean wheel diameter (D12) at the outlet edge (13) of each blade (12) of the compressor wheel (37) is less than 88 percent of the maximum outer diameter (D2) of the blade (2) of the turbine wheel 1.

11. An exhaust gas turbocharger, according to claim 10, in which the compressor wheel (37) is diagonally traversed from the center to the periphery of said wheel.

12. An exhaust gas turbocharger according to claim 10, in which the outlet edges (13) of the blades (12) of the compressor wheel (37), projected into meridian planes, are substantially perpendicular to the direction (38) of the flow of air passage projected into the same meridian planes.

13. An exhaust gas turbocharger, according to claim 10, in which the cross sections of the blades (12) of the compressor wheel (37) are arranged radially in sectional planes perpendicular to the wheel axis.

14. An exhaust gas turbocharger according to claim 10, in which each of the blades (12) of the compressor wheel (37) forms at the respective outlet edge (13) an outlet angle .beta.2 of less than 90.degree..

15. An exhaust gas turbocharger, according to claim 10, in which the mean outlet diameter (D12) of the compressor wheel (37) is at most equal to the outer diameter (D11) of the compressor inlet edge (26).

16. An exhaust gas turbocharger, according to claim 1, in which each outlet edge of said compressor wheel and each inlet edge of said turbine wheel projects axially inwardly beyond the axially inner wall of the respective wheel.