Method of reducing turbine wheel high cycle fatigue in sector-divided dual volute turbochargers

Abstract

High cycle fatigue (HCF) in a turbine wheel of a sector-divided dual volute turbocharger, particularly a turbocharger where the tongue-to-blade gap is as small as from 1-3% of the wheel diameter, is reduced using a turbine wheel with (blade stiffness/backwall stiffness100) between 41 and 44.

Claims

1. A turbine of an exhaust-gas turbocharger, comprising: a sector-divided dual volute turbine housing (4) comprising first and second spiral-shaped exhaust-gas channels (5, 6) narrowing circumferentially inwardly, delimited radially to the outside by an inner surface of the turbine housing (4) and separated from each other by first and second tongues (16, 17), each tongue associated with one exhaust-gas channel, a turbine wheel (2) mounted to be rotatable about an axis of rotation within the turbine housing (4) and comprising a hub having a wheel backwall (21) which is of closed configuration as viewed from an axial direction and which has a circular circumference, a hub which extends from the wheel backwall and which tapers along the axial direction, and a multiplicity of turbine blades (20) which extend from the wheel backwall (21) and from the hub and which extend radially outward, wherein each of said multiplicity of turbine blades (20) has a stiffness, wherein the backwall (21) has a stiffness, and wherein a stiffness ratio (blade stiffness/backwall stiffness100) is from 41 to 44.

2. The turbine of claim 1, wherein the stiffness ratio is from 41.2 to 43.5.

3. The turbine of claim 1, wherein the stiffness ratio is from 41.5 and 43.3.

4. The turbine of claim 1, wherein a gap is provided between the tongues (16, 17) and passing turbine blades (20), wherein the gap is from 1-7% of the wheel diameter.

5. The turbine of claim 1, wherein a gap is provided between the first and second tongues (16, 17) and passing turbine blades (20), wherein the gap is from 1-3% of the wheel diameter.

6. The turbine of claim 1, wherein a gap is provided between the tongues (16, 17) and passing turbine blades (20), wherein the gap is from 1-2% of the wheel diameter.

7. An exhaust gas turbocharging system (1) comprising a first exhaust manifold (9) and a second exhaust manifold (10), a pressure-charged combustion engine (8) having at least one cylinder head comprising at least two cylinders (C.sub.1, C.sub.2), in which each cylinder has at least one outlet port for discharging the exhaust gases from the cylinder, wherein at least two cylinders are configured in such a way that they form first and second cylinder groups, each cylinder group comprising at least one cylinder, the first group supplying exhaust to said first exhaust manifold (9) and the second cylinder group supplying exhaust gas to said second exhaust manifold (10), a sector-divided dual volute turbine housing (4) comprising first and second spiral-shaped exhaust-gas channels (5, 6) narrowing circumferentially inwardly, delimited radially to the outside by an inner surface of the turbine housing (4) and separated from each other by first and second tongues (16, 17), each tongue associated with one exhaust-gas channel (5, 6), the first spiral-shaped exhaust-gas channel (5) receiving exhaust flow from the first exhaust manifold (9) and the second spiral-shaped exhaust-gas channel (5) receiving exhaust flow from the second exhaust manifold (10), a turbine wheel (2) mounted to be rotatable within the turbine housing (4) comprising a hub having a wheel backwall (21) which is of closed configuration as viewed from an axial direction and which has a circular circumference, a hub which extends from the wheel backwall and which tapers along the axial direction, and a multiplicity of turbine blades (20) which extend from the wheel backwall and from the hub and which extend radially outward, wherein each blade has a stiffness, wherein the backwall has a stiffness, and wherein stiffness ratio (blade stiffness/backwall stiffness100) is from 41 to 44.

8. The exhaust gas turbocharging system of claim 7, wherein the stiffness ratio is from 41.2 to 43.5.

9. The exhaust gas turbocharging system of claim 7, wherein the stiffness ratio is from 41.5 and 43.3.

10. The exhaust gas turbocharging system of claim 7, wherein a gap is provided between the tongues (16, 17) and passing turbine blades (20), wherein the gap is from 1-7% of the wheel diameter.

11. The exhaust gas turbocharging system of claim 7, wherein a gap is provided between the tongues (16, 17) and passing turbine blades (20), wherein the gap is from 1-3% of the wheel diameter.

12. The exhaust gas turbocharging system of claim 7, wherein a gap is provided between the tongues (16, 17) and passing turbine blades (20), wherein the gap is from 1-2% of the wheel diameter.

Description

BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS

(1) FIG. 1 shows an in-line four cylinder internal combustion engine with exhaust flow separation maintained through exhaust manifolds to a sector-divided dual volute turbine housing;

(2) FIG. 2 shows an alternative sector-divided dual volute turbine;

(3) FIG. 3 shows a turbine wheel with an exaggerated vibration at the blade leading edge and how blade bending is measured;

(4) FIG. 4 shows a turbine wheel with an exaggerated vibration at the backwall (disk) and how backwall bending is to be measured; and

(5) FIGS. 5, 6 show turbine wheel blade failure.

DETAILED DESCRIPTION OF THE INVENTION

(6) As diagrammatically illustrated in FIGS. 1 and 2, the invention relates to reduction of HCF in a turbine wheel (2) in a sector-divided turbine housing (4) for an exhaust-gas turbocharger system (1) as illustrated in FIG. 1. In this illustrative embodiment, the engine (8) has four cylinders (C.sub.1, C.sub.2, C.sub.3, C.sub.4), which release pulses of exhaust gas into a dual exhaust manifold (7). In a typical in line four cylinder engine, the firing order is 1-3-4-2, meaning that the first cylinder C.sub.1 combusts first, then the third cylinder C.sub.3 combusts second, then the fourth cylinder C.sub.4 combusts third, and the second cylinder C.sub.2 combusts fourth. Two cylinders (C.sub.1, C.sub.4) are coupled to a first exhaust manifold pipe (9), and two cylinders (C.sub.2, C.sub.3) are coupled to a second exhaust manifold (10). The exhaust manifolds have flanges (11, 13) via which they are coupled to turbocharger turbine housing flange (13) with maintenance of continuity of separate exhaust flow channels and thus separation of exhaust pulses (15). By coupling the first and fourth cylinders (C.sub.1, C.sub.4) together and coupling the second and third cylinders (C.sub.2, C.sub.3) together, the exhaust pulses released from each cylinder are separated by a maximum amount for the configuration. Stated differently, after the exhaust pulse from the first cylinder C.sub.1 enters the first exhaust manifold pipe (9), there is a pause before the exhaust pulse from the fourth cylinder C.sub.4 enters the first exhaust manifold (9) because the exhaust pulse from the third cylinder C.sub.3 first enters the second exhaust manifold (10). Likewise, after the exhaust pulse from the fourth cylinder C.sub.4 enters the first exhaust manifold pipe (9), there is a pause before the exhaust pulse from the first cylinder C.sub.1 enters the first exhaust manifold (9) because the exhaust pulse from the second cylinder C.sub.2 first enters the second exhaust manifold (10). These pauses result in separation between pulse energies in the exhaust manifolds (9, 10) such that interference between adjacent exhaust pulses in a manifold is reduced. The pulses from the third and fourth cylinders (C.sub.3, C.sub.4) are similarly separated. The turbine housing (4) has at least two symmetric or asymmetric flow channels (5, 6) that are fluidically substantially separate and can be flowed through by the exhaust gas. A turbine wheel (2) that is rotatably accommodated in the turbine housing (4) can be driven by the exhaust gas impinging on the periphery (14) of the turbine wheel.

(7) The multiple volutes (5, 6) channel the exhaust gas so that it is introduced generally evenly to the circumference (14) of the turbine wheel (2) and expands in the turbine wheel whereby the turbine wheel may convert the exhaust gas into rotational mechanical energy. The pressure and temperature drop is converted into kinetic energy.

(8) From an energy extraction perspective, it is desirable to keep the gap between tongues (16, 17) and turbine wheel as small as possible. However, as each blade of the rotating turbine wheel passes close to the flow obstruction formed by the tongues (16, 17), pressure variations impart an excitation to the blade, causing the blade to vibrate, in the way a tuning fork will vibrate when struck. As the tongues (16, 17) are positioned closer to the turbine wheel, the pressure variations increase, resulting in a greater excitation and therefore greater vibration of the blade. Moreover, with two tongues, each blade will experience two of these excitations per revolution. To minimize this energy transfer to the blades, it is conventional that the inner boundary of the volute (be it meridionally divided or sector divided) has a minimum diameter not less than 106% of the turbine wheel diameter. In contrast, in accordance with the present invention, when the turbine wheel possesses the specified stiffness ratio, it is possible to reduce HCF to acceptable values even when the tongue-to-blade gap is as small as from 1-3% of the wheel diameter, preferably 1-2%.

(9) As illustrated in FIG. 3, in such a turbocharger with sector-divided dual volute turbine housing, when aerodynamic forces act on the turbine blades, the blades are deflected from their static shape, being bent backwards and forwards. The accumulation of vibrational energy without an adequate dissipation mechanism can lead to increasing amplitude of vibration. Repeated bending or deflection leads to material fatigue, cracking and an ultimate fracture, referred to as high cycle fatigue (HCF) failures.

(10) As illustrated in FIG. 4, vibration of the wheel can also take place at the turbine wheel backwall. Vibration of the backwall can also lead to turbine wheel failure, with backwall failure usually being more serious than blade failure.

(11) During operation, the blades can be subject to vibratory excitation at frequencies which coincide with integer multiples, referred to as harmonics, of the turbine wheel's rotational frequency. As a result of the vibratory excitation, the blades can undergo vibratory deflections that create vibratory stress on the blades. If the vibratory excitation occurs in an expected operating speed range of the turbine, the vibratory stresses can create high cycle fatigue and cracks over time.

(12) Examples of blade failure are illustrated in FIGS. 5 and 6.

(13) HCF is characterized by low amplitude high frequency elastic strains. HCF is a type of fatigue caused by small elastic strains under a high number of cycles before failure occurs.

(14) Part of the physical design of the turbine housing is the volute, the function of which is to control the inlet conditions to the turbine wheel such that the inlet flow conditions provide the most efficient transfer of power from the energy in the exhaust gas to the power developed by the turbine wheel, combined with the best transient response characteristics. Theoretically the incoming exhaust flow from the engine is delivered in a uniform manner from the volute to a vortex centered on the turbine wheel axis. To do this, the cross-sectional area of the volute is at a maximum perpendicular to the direction of flow gradually and continuously decreasing until it becomes zero. The inner boundary of the dual volute can describe a spiral, of minimum diameter from 102% to 106% of the turbine wheel diameter (i.e., wheel-to-tongue gap from 1-3% of the wheel diameter).

(15) The turbine and turbine wheels according to the present invention are specifically designed to harness pulse energy and convert it to rotational velocity. Thus, the conversion of pressure and velocity from the exhaust gas for a pulse flow turbine wheel in a divided turbine housing is greater than the conversion of pressure and velocity from a steady state exhaust flow to the turbine wheel velocity.

(16) The operation of the turbocharger system of FIG. 1 will now be explained. In sector-divided dual volute (two-scroll, two-tongue) turbo systems, divided turbo manifolds have been designed to feed divided chambers in the volute to enhance the benefits of pulse energy by separating exhaust gas flow into two branches. Gas flow from certain cylinders of an in-line four-cylinder engine, such as C.sub.2 and C.sub.3, pass through one branch (passageway) of the manifold, and gas from other cylinders, such as C.sub.1 and C.sub.4, pass through a separate branch. Gas flow from each branch from respective cylinders stay divided in the twin-scroll in the volute of the turbine housing. The resulting two feed ports deliver opposite and equal firing pulsations (15) to improve turbine efficiency and reduce manifold complexity. Divided manifold runners (9, 10), such as for in-line four cylinder configurations, enhance pulse utilization by maintaining separation of alternating pulses (15) of exhaust flow. Similarly, a six-cylinder configuration may have cylinders 1, 2 and 3 combined and have cylinders 4, 5 and 6 combined as separate branches into two feed ports providing alternating pulses. Twin-scroll turbo systems may have higher backpressure at low rpm (which may help turbo spool-up) and lower backpressure at high rpm (which may help top-end performance).

Example

(17) The physical attributes of a large number of turbine wheels were carefully measured.

(18) Turbine wheel blade stiffness is measured by applying a force of 100 Newtons to the blade where it is most liable to bend, namely normal (N.sub.B; FIG. 3) to the radially outer circumferential end of the trailing edge of the blade. Loading could be virtual in the case of computer simulation, or by load cells, etc., for physical turbine wheels.

(19) Turbine wheel backwall stiffness is measured by applying a force of 100 Newtons to the wheel backwall where it is most liable to bend, namely normal (N.sub.BW; FIG. 3) to the outer circumferential edge of the backwall disk half-way between blades, and measuring deflection.

(20) Blade thickness, blade uniformity, wheel materials, turbine housing geometry, hub-to-blade-tip variation, wheel balance and eccentricity, tongue-to-blade gap distance, blade-to-blade frequency variations, etc were also measured.

(21) The turbine wheels were subjected to prolonged operation, either by computer simulation or bench top testing, and results carefully measured.

(22) As a result of analysis of a large pool of data, it was surprisingly discovered that there was a low yet commercially-acceptable level of HCF in turbine wheels when the blade to backwall stiffness ratio is between 41 and 44, more preferably when the stiffness ratio is between 41.2 and 43.5, most preferably when the stiffness ratio is between 41.5 and 43.3. The units used in the stiffness ratio may be Newton/mm (i.e., force/displacement), for example.

(23) More specifically, the lower bounds of acceptable stiffness ratio are found in a turbine wheel having a blade deflection of 0.02473 mm at 100 Newton and a blade stiffness of 4042.2006, and a disk deflection of 0.010278 mm at 100 Newton and disk stiffness of 9729.5194. The blade to disk stiffness ratio, multiplied by 100, is 41.545737.

(24) The upper bounds of the acceptable stiffness ratio are found in a turbine wheel having a blade deflection of 0.0178 mm at 100 Newton and a blade stiffness of 5620, and a disk deflection of 0.0769 mm at 100 Newton and disk stiffness of 13000. The blade to disk stiffness ratio, multiplied by 100, is 43.23921.

(25) Accordingly, it was discovered that HCF is reduced to acceptable levels when blade to backwall stiffness ratio (100) is between 41 and 44, more preferably when the stiffness ratio is between 41.2 and 43.5, most preferably when the stiffness ratio is between 41.5 and 43.3.

(26) It will be appreciated that the configurations and routines disclosed herein are exemplary in nature, and that these specific embodiments are not to be considered in a limiting sense, because numerous variations are possible. For example, the above technology can be applied to I-4, I-6, V-6, V-12, opposed 4, and other engine types. Moreover, the above-described turbine housing could be used with a two cylinder engine, such as a V-twin or inline two-cylinder engine, wherein exhaust gas pulses from each of the two cylinders is directed to one of the two volutes.

(27) In at least one embodiment, the above-described turbine configuration may be used with a turbine housing with a volute that is both sector divided and meridonally divided. Stated differently, the turbine housing may be divided into two volutes with circumferentially-spaced tongues and each of the two volutes may be divided transversely by an integral wall. Such a turbine housing would have four divided volutes. Such a turbine housing is described by U.S. Pat. No. 3,218,029 and U.S. Publication No. 2017/0183975 A1. Such a turbine housing may be advantageously applied with a four cylinder engine such that the exhaust pulses for each cylinder travels through a separate volute. Such a turbine housing could also be used with an eight cylinder engine, a twelve cylinder engine, or a sixteen cylinder engine (or any other engine with a total number of cylinders that is an integer multiple of four).

(28) The above-described turbine could also be used with an internal combustion engine having an uneven number of cylinders, such as a three-cylinder engine or a five-cylinder engine. In such configurations, a greater number of cylinders would exhaust to a first volute than to a second volute. For example, in a three-cylinder engine, cylinders 1 and 2 may exhaust to the first volute (5) and cylinder 3 may exhaust to the second volute (6). Likewise, in a five-cylinder engine, cylinders 1, 3, and 4 may exhaust to the first volute (5) and cylinders 2 and 5 may exhaust to the second volute (6). To accommodate the different volumetric flow rates caused by the different numbers of cylinders exhausting to the two volutes, the two volutes (5, 6) may have different cross-sectional flow areas. In the above-described examples of three and five-cylinder engines, the second volute (6) may have a smaller cross-sectional area than the first volute (5) since less exhaust gas passes through the second volute (6).

(29) All of the above U.S. patents, U.S. patent application publications, U.S. patent applications, foreign patents, foreign patent applications and non-patent publications referred to in this specification, are incorporated herein by reference in their entirety.

(30) While particular elements, embodiments and applications of the present invention have been shown and described, it will be understood, of course, that the invention is not limited thereto since modifications may be made by those skilled in the art without departing from the spirit and scope of the present disclosure, particularly in light of the foregoing teachings.