Turbine wheel, a turbine and use thereof

Abstract

A turbine wheel for low-pressure ratio applications is disclosed. The ratio of the outlet area of the wheel (A2) to the inlet area of the wheel (A1) is less than approximately 0.4. In an embodiment, the wheel is a radial or mixed-flow wheel.

Claims

1. An exhaust system for a vehicle comprising: a turbocharger; and a turbocompound unit comprising a turbine downstream of the turbocharger, wherein the turbine is configured to operate with an inlet to outlet pressure ratio between 1.02 and 1.2 and comprises: a turbine wheel for low pressure ratio applications, wherein the ratio of the outlet area of the wheel (A2), the outlet area defined as the area described by rotating a first edge of a blade of the turbine wheel about the axis, the first edge being an edge arranged to be adjacent an outlet, to the inlet area of the wheel (A1), the inlet area defined as the area described by rotating a second edge of the blade of the turbine wheel about the axis, the second edge being an edge arranged to be adjacent an inlet, is between 0.3 and 0.4, wherein the ratio of the radius of the root of the blades adjacent the outlet to the radius of the tip of the blades adjacent the outlet is between 0.2 and 0.7.

2. The exhaust system according to claim 1, wherein a ratio of the radius of the tip of the blades adjacent the outlet to the radius of the tip of the blades adjacent the inlet is less than 1.0.

3. The exhaust system according to claim 1, wherein a ratio of the radius of the tip of the blades adjacent the outlet to the radius of the tip of the blades adjacent the inlet is between 0.6 and 0.9.

4. The exhaust system according to claim 1, wherein the exit relative flow angle is less than 55 degrees.

5. The exhaust system according to claim 1, wherein the exit relative flow angle is between 41 degrees and 55 degrees.

6. The exhaust system according to claim 1 and further being a radial-flow turbine wheel.

7. The exhaust system according to claim 1 and further being a mixed-flow turbine wheel.

8. The exhaust system according to claim 1 and further comprising a shroud at least partly covering the turbine wheel to define an inlet and an outlet of the turbine.

9. The exhaust system according to claim 1, wherein a ratio of a radius (R3) of a hub of the turbine wheel adjacent the outlet to a radius (R4) of the outlet defined by the shroud is less than 0.7.

10. The exhaust system according to claim 1, wherein a ratio of a radius (R4) of the outlet defined by the shroud to a radius (R1) of the inlet defined by the shroud is less than 1.0.

11. A turbine according to claim 8, wherein a ratio of a radius (R3) of a hub of the turbine wheel adjacent the outlet to a radius (R4) of the outlet defined by the shroud is between 0.2 and 0.7.

12. A turbine according to claim 8 wherein a ratio of a radius (R4) of the outlet defined by the shroud to a radius (R1) of the inlet defined by the shroud is between 0.6 and 0.9.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

(1) Specific embodiments of the invention will be described below by way of example only and with reference to the accompanying drawings, in which:

(2) FIG. 1 shows an existing turbocharger design.

(3) FIG. 2 shows a typical arrangement of a turbocharged engine.

(4) FIG. 3 is a turbine map showing the normalised total-to-static efficiency (vertical axis) vs. Pressure ratio (PR) (horizontal axis). The total-to-static efficiency curves are plotted for constant speed lines as indicated in the legend by the Speed Parameter (SP) given in terms of equivalent percentage speed. This figure gives a comparison between the normalised total-to-static efficiency obtained with prior art applications and that obtained with the embodiments of the present invention (LPT Design).

(5) FIG. 4 is a chart correlating the blade loading coefficient () (vertical axis) and the flow coefficient () (horizontal axis) with the turbine total-to-static efficiency (dashed lines).

(6) FIG. 5 is an axial view of a turbine wheel that embodies the invention, and shows also a flow velocity triangle at the inlet to the turbine wheel (1). In this Figure are shown the absolute flow velocity (C1), the relative flow velocity (W1), the peripheral speed (U1), the absolute flow angle (1) and the relative flow angle (1);

(7) FIG. 6 shows the sensitivity of absolute flow angle (1) (horizontal axis) and the normalised turbine total-to-static efficiency (vertical axis).

(8) FIG. 7 is a radial view of a turbine wheel that embodies the invention, and shows also a flow velocity triangle at the exit to the turbine wheel (2). In this Figure are shown the absolute flow velocity (C2), the relative flow velocity (W2), the peripheral speed (U2), the absolute flow angle (2) and the relative flow angle (2);

(9) FIG. 8 shows the sensitivity of exit relative flow angle (2) (horizontal axis) and the normalised turbine total-to-static efficiency (vertical axis).

(10) FIG. 9 shows blade profile obtained as a projection on the longitudinal plane.

(11) FIG. 10 shows the sensitivity of the exit relative flow angle (2) (horizontal axis) with the ratio between the exit radius (R4) and the inlet radius (R.sub.1) & the ratio between the exit hub radius (R3) and the exit shroud radius (R4) (vertical axis).

(12) FIG. 11 is an isometric view of a turbine wheel that embodies the invention: the inlet (A1) and the exit (A2) area to the turbine which have been considered in the design are indicated by the dashed areas.

(13) FIG. 12 shows the sensitivity of exit relative flow angle (2) (horizontal axis) with the ratio between the exit area (A2) and the inlet area (A1) to the turbine (A2/A1) (vertical axis).

(14) FIG. 13 shows the sensitivity of the ratio between the exit area (A2) and the inlet area (A1) to the turbine (A2/A1) (horizontal axis) with the normalised turbine total-to-static efficiency.

(15) FIG. 14 shows the difference between a radial and mixed-flow turbine.

SPECIFIC DESCRIPTION OF CERTAIN EXAMPLE EMBODIMENTS

(16) The description of the design of a low pressure turbine will now be undertaken. The non-dimensional design procedure is intended to determine the overall turbine configuration.

(17) Embodiments of the invention are described with reference to FIGS. 4 to 14.

(18) The configuration of a turbine is started with two parameters, the blade loading coefficient and the flow coefficient . The blade loading and the flow coefficient are two non-dimensional parameters; is defined as the ratio between the actual enthalpy changes (U2.Math.C2.Math.tan 2U1.Math.C1 tan 1) and the square of the peripheral speed (U1), while is defined as the ratio between the meridional component of the absolute flow velocity (CM1) and the peripheral speed (U1). The blade loading and the flow coefficient are uniquely correlated to the total-to-static efficiency as shown in FIG. 4.

(19) FIG. 4 shows that the optimum total-to-static efficiency region falls in the range of 0.1 to 0.3 for the flow coefficient () and 0.7 to 1.1 for the blade loading coefficient ().

(20) This constrains the values of the absolute flow angle (1) (FIG. 5) to have values below approximately 80. This amounts to a first requirement.

(21) This requirement is shown in FIG. 6 where the total-to-static efficiency is plotted against the absolute flow angle 1. The figure shows that the total-to-static efficiency increases as al increases. However values too high for 1 cannot be selected as it would cause the absolute flow velocity (C1) to be tangential and it would cause high incidence loss. This will be referred to as a second requirement.

(22) The requirements set out above constrain the number of blades to vary between 8 and 13. This ensures manufacturability and avoids blade crowding at the exit to the turbine.

(23) All preceding requirements must be satisfied in a low pressure ratio condition (PR1.02-1.2) which constrains the wheel geometry to be different from prior art applications of a micro radial/mixed turbine.

(24) Further turbine development is carried out by evaluating the rotor discharge condition (FIG. 7). This is determined by varying the exit relative flow angle (2) (horizontal axis) with respect to the turbine total-to-static efficiency (vertical axis) as shown in FIG. 8.

(25) From FIG. 8 it can be seen that the total-to-static efficiency increases as the exit relative flow angle 2 increases. Thus the value of 2 should be set as high as possible. However a large 2 would increase the amount of flow separation and secondary flows which contribute to total-to-static efficiency loss, thus further limiting the operating range of the turbine.

(26) An optimum exit relative flow angle (2) therefore needs to be defined in order to prevent flow separation and recirculation to occur but still maintaining higher total-to-static efficiency.

(27) The selection of 2 has a direct impact on the rotor wheel geometry. The geometric parameters which define that geometry are given in FIG. 9.

(28) In this figure are shown the radiuses at the leading edge (R1 and R2) and the trailing edge (R3 and R4) of the turbine wheel: R1: rotor shroud diameter (leading edge) R2: rotor hub diameter (leading edge) R3: rotor hub diameter (trailing edge) R4: rotor shroud diameter (trailing edge)

(29) The correlation between the exit relative flow angle 2 and the wheel geometry is shown in FIG. 10 where the ratio between the hub exit radius (R3) and the shroud exit radius (R4) is determined for different exit relative flow angles (2). FIG. 10 shows that the radius ratio R3/R4 increases as the exit relative flow angle (2) and this would correspond to an increase in total-to-static efficiency (FIG. 7).

(30) The radius ratio R3/R4 must be retained to values ranging within 0.2 and 0.7: values of R3/R4 less than 0.2 would limit the strength of the shaft while values of R3/R4>0.7 would correspond to large hub thus increasing the inertia of the wheel.

(31) The selection of 2 and R3/R4 as set out above also defines the exit to inlet conditions of the turbine blade. The ratio between the shroud exit radius (R4) and the shroud inlet radius (R1) is evaluated and plotted against the exit relative flow angle (2), FIG. 10.

(32) FIG. 10 shows that the radius ratio R4/R1 varies linearly with 2 and cannot exceed 1.0 since it would cause an expansion too large through the wheel. Hence the radius ratio R4/R1 has to vary between 0.6 and 0.9.

(33) In order to satisfy the low pressure ratio condition whilst still maintaining high total-to-static efficiency, the requirements set out hereinabove can be obtained by retaining a low value of the ratio between the exit area (A2) and the inlet area (A1), FIG. 11.

(34) FIG. 12 shows the variation of the area ratio (A2/A1) (vertical axis) with the exit relative flow angle (2). This figure shows that in order to meet the required flow conditions for 2, a low value of the area ratio must be maintained. This condition is directly related with the turbine total-to-static efficiency, as shown in FIG. 13. The figure shows that an increase in A2/A1 leads to an increase in the total-to-static efficiency.

(35) As a consequence of the direct correlation between the exit relative flow angle 2 and the area ratio A2/A1, the maximum total-to-static efficiency conditions are obtained for A2/A1 lower than 0.4.

(36) The requirements set out hereinabove fix the blade geometry for a radial or mixed-flow turbine wheel operating at low pressure conditions. The shroud inlet radius (R1), the inlet shroud exit radius (R4), the hub exit radius (R3), the exit relative flow angle (2) and the area ratio condition (A2/A1) uniquely define the blade geometry.

(37) Once the hub and shroud geometrical results had been defined a standard 4.sup.th degree Bezier polynomial curve is used to define the blade profiles starting from the hub up to the shroud and to generate single camber-line curves.

(38) The blade geometry is finally completed by using a radial fibre blade design method. The distinction between a radial turbine and a mixed-flow turbine is the cone angle () at the inlet to the turbine (FIG. 14). By definition a radial turbine has an inlet blade angle B=0 and the blade radial fibre requirement constrains the cone angle to be fixed at =90. In a mixed flow turbine the zero blade angle limitation can be overcome by radially sweeping the inlet blade of a radial turbine but still maintaining the radial fibre condition (90 and B0. In addition to this, in a radial turbine the shroud inlet radius (R1) is equal to the hub inlet radius (R2), R1=R2. However the procedure remains unaltered independently of whether a radial or a mixed-flow turbine is designed.

(39) It will be appreciated that the approach disclosed herein of adapting the ratio A2/A1 such that it is below approximately 0.4 is in contrast to established approaches to varying turbine performance. Specifically, it will be understood that established approaches teach the reshaping of the profile of the turbine wheel and of the turbine shroud, and have not hitherto considered the ratio A2/A1 or indeed modifying the turbine such that this ratio is below approximately 0.4 to give a turbine that is especially suited to low pressure ratio applications.

(40) The current disclosure applies both to radial and mixed-flow turbines.