Centrifugal compressor with diffuser with throat

Abstract

A diffuser is proposed which is formed as the gap between rotationally-symmetric surfaces which face each other. Moving in the radial direction, the axial extent of the gap generally decreases to a minimum value in a throat portion of the diffuser, and then generally increases again. The distance from the rotational axis of the compressor to the throat may be approximately at least 125% of the radius of the compressor wheel. The inventors have found that a throat at this distance from the rotational axis may lead to higher efficiency at high flow rates, especially for relatively low turbo speeds. This is because the spacing between the compressor wheel and the throat permits diffusion of the gas streams leaving the compressor wheel.

Claims

1. A compressor for a turbomachine, the compressor comprising: a housing defining an inlet, an outlet and a compressor chamber; a compressor wheel mounted within the compressor chamber for rotation about a rotational axis, the compressor wheel having a plurality of blades; the housing defining: a scroll radially outward of the compressor chamber and communicating with the outlet of the housing; and a diffuser space between an radially-extending shroud surface of the housing and a radially-extending hub surface, the diffuser space having an inlet communicating with the compression chamber and an outlet into the scroll, the diffuser space being rotationally symmetric about the axis, the diffuser space having: a throat portion where the diffuser has minimum axial extent; a radially-inner portion extending radially-inwardly from the throat portion, and throughout which the diffuser space has a greater axial extent than said minimum axial extent; and a radially-outer portion extending radially-outwardly from the throat portion to the scroll, and throughout which the diffuser space has a greater axial extent than said minimum axial extent; the radially-outer edge of the radially-inner portion of the diffuser space being at a radial distance from the rotational axis which is no less than 125% of the radius of the compressor wheel; and the radially-inner edge of the radially-outer portion of the diffuser space being at a radial distance from the rotational axis which is no more than 140% of the radius of the compressor wheel.

2. A compressor according to claim 1 in which the radially-outer edge of the radially-inner portion of the diffuser space is at a distance from the rotation axis which is no less than 130% of the radius of the compressor wheel.

3. A compressor according to claim 1 in which the radially-outer edge of the radially-inner portion of the diffuser space is at a distance from the rotation axis which is no less than 140% of the radius of the compressor wheel.

4. A compressor according to claim 1 in which, at the radially-outer edge of the radially-outer portion of the diffuser space, the hub surface has a tangent perpendicular to the circumferential direction, which is at an angle of less than 90 degrees to the axial direction.

5. A compressor according to claim 1 in which the ratio of the distance from the rotational axis to the radially-inner edge of the radially-outer portion, to the distance from the rotational axis to the radially-outer edge of the radially-outer portion is in the range 75% to 90%.

6. A compressor according to claim 1 in which the axial extent of the diffuser space at the throat position is at least 65% of the axial extent of the blades at their radially-outer ends.

7. A compressor according to claim 1 in which the axial extent of the diffusion space increases in the radially-outward direction throughout the radially-outer portion.

8. A compressor according to claim 1 in which the axial extent of the diffusion space increases in the radially-inner direction throughout the radially-inner portion, and the radially-inner portion extends inwardly to a position which is spaced from the rotational axis by at most 110% of the radius of the compressor wheel.

9. A compressor according to claim 1 in which the throat portion has no radial extent.

10. A compressor according to claim 1 in which, between the compressor wheel and the scroll, the shroud wall is non-concave as viewed in a plane including the axis.

11. A compressor according to claim 10 in which, between the compressor wheel and the scroll, the shroud wall is convex as viewed in a plane including the axis.

12. A compressor according to claim 1 in which, at the radially-outer edge of the radially-outer portion of the diffuser space, the hub surface has a tangent perpendicular to the circumferential direction which is at an angle of no more than 80 degrees to the axial direction.

13. A compressor according to claim 1 in which the ratio of the distance from the rotational axis to the radially-inner edge of the radially-outer portion, to the distance from the rotational axis to the radially-outer edge of the radially-outer portion is less than 85%.

14. A turbocharger including a compressor, the compressor comprising: a housing defining an inlet, an outlet and a compressor chamber; a compressor wheel mounted within the compressor chamber for rotation about a rotational axis, the compressor wheel having a plurality of blades; the housing defining: a scroll radially outward of the compressor chamber and communicating with the outlet of the housing; and a diffuser space between an radially-extending shroud surface of the housing and a radially-extending hub surface, the diffuser space having an inlet communicating with the compression chamber and an outlet into the scroll, the diffuser space being rotationally symmetric about the axis, the diffuser space having: a throat portion where the diffuser has minimum axial extent; a radially-inner portion extending radially-inwardly from the throat portion, and throughout which the diffuser space has a greater axial extent than said minimum axial extent; and a radially-outer portion extending radially-outwardly from the throat portion to the scroll, and throughout which the diffuser space has a greater axial extent than said minimum axial extent; the radially-outer edge of the radially-inner portion of the diffuser space being at a radial distance from the rotational axis which is no less than 125% of the radius of the compressor wheel; and the radially-inner edge of the radially-outer portion of the diffuser space being at a radial distance from the rotational axis which is no more than 140% of the radius of the compressor wheel.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

(1) A non-limiting embodiment of the invention will now be described, for the sake of example only, with reference to the following figures, in which:

(2) FIG. 1 is a cross-sectional drawing of a known turbocharger;

(3) FIG. 2 shows a baseline configuration for a diffuser;

(4) FIG. 3 shows schematically a configuration of a diffuser which is an embodiment of the invention;

(5) FIG. 4 shows the configuration of three embodiments of the invention compared to the baseline configuration;

(6) FIG. 5 is composed of FIG. 5(a) which shows the pressure ratio (inlet to outlet), and FIG. 5(b) which shows the efficiency, as a function of the mass flow for the first of the embodiments;

(7) FIG. 6 is composed of FIG. 6(a) which shows the pressure ratio (inlet to outlet), and FIG. 6(b) which shows the efficiency, as a function of the mass flow for the second of the embodiments; and

(8) FIG. 7 is composed of FIG. 7(a) which shows the pressure ratio (inlet to outlet), and FIG. 7(b) which shows the efficiency, as a function of the mass flow for the third of the embodiments.

DETAILED DESCRIPTION OF THE EMBODIMENTS

(9) Referring firstly to FIG. 2, a baseline configuration is shown for the diffuser 39 of the turbocharger of FIG. 1. The baseline configuration is a comparative example used below in computational simulation comparisons with embodiments of the invention.

(10) Eight radially-spaced reference positions in the diffuser are marked 1-8 in FIG. 2. Table 1 shows the radial position of these reference positions, measured from the centre of the rotational axis of the shaft 18. The radial position of the radially outer tip of the blades of the compressor wheel 16 (not shown) is denoted as 41, and is at a distance 54 mm from the rotational axis of the shaft 18.

(11) TABLE-US-00001 TABLE 1 Reference Radial distance of reference point from the next Radial distance from the axis of reference point in Reference the shaft 18 the radially-inward position mm % of wheel diameter direction (mm) 1 57.5 106.481 — 2 62.45 115.648 4.95 3 67.4 124.815 4.95 4 72.35 133.981 4.95 5 77.3 143.148 4.95 6 82.25 152.315 4.95 7 87.25 161.574 5 8 92.86 171.963 5.61

(12) The reference position 1 of the diffuser of the baseline configuration has a first axial width b.sub.2. The diffuser 39 becomes narrower linearly at successive positions in the radially-outward direction, until reference position 2. Then it has substantially constant width until the outlet reference position 8. At the reference position 1, the angle between the tangent to the hub surface 20 (perpendicular to the circumferential direction) and the axial direction is marked as a.sub.2. At the outlet position 8, the angle between the tangent to the hub surface (measured in a plane including the rotational axis) and the axial direction is marked as a.sub.3, and the axial width at the outlet 8 is denoted by b.sub.3.

(13) By contrast, FIG. 3 shows schematically the shape of the diffuser in certain embodiments of the invention. Distances in FIG. 3 are not drawn to scale, and below we supply distance parameters defining three specific embodiments. In each case, the diffuser is rotationally symmetric about the axis of the shaft, and the reference positions 1 to 8 are in the same radial positions as in the baseline configuration shown in FIG. 2.

(14) The diffuser gap has a narrowest axial extent at a single, radial position 44, referred to as the throat position. The portion of the diffuser which is radially-inward from the throat portion 44 is the radially-inner portion 42. The portion of the diffuser which is radially-outward from the throat portion 44, and extends to the scroll, is the radially-outer portion 43. The radially-inner portion 42 and radially-outer portion 43 of the gap touch at the throat position 44 because the throat position 44 has no radial extent.

(15) However, more generally, there may be a range of radial positions at which the gap has the same, minimal axial extent. In other words, the diffuser has a throat portion which may have any radial extent. Throughout the throat portion, all positions on the shroud surface 20 are axially spaced by this same axial distance from respective positions on the hub surface 21. The throat portion spaces the radially-inner portion of the diffuser radially from the radially-outer portion.

(16) The arrangement of FIG. 3 may be considered as a limiting case of this, in which the throat portion has zero radial extent: the portion of the shroud surface 20 which is closest to the hub surface 21 is just a circular line at the throat position 44. In other words, in the arrangement of FIG. 3, the throat portion of the gap is the single, radial throat position 44.

(17) We now turn to more precise definitions of the parameters of the baseline configuration of FIG. 2, and the three embodiments with the general shape shown schematically in FIG. 3.

(18) As in the baseline configuration, in all three embodiments the compressor wheel 16 has a diameter of 108 mm, i.e. a radius of 54 mm. Table 2 shows further parameters which are in common between the baseline configuration and the three embodiments. The impeller tip width means the axial length of the blades of the compressor wheel 6 at their radially-outer point. The radially-outer edge of the blade has equal distance from the rotational axis along the whole length of the blade. As mentioned above, the diffuser inlet width b.sub.2 is the axial width of the diffuser at the reference position 1. The diffuser length is the radial distance from the reference position 1 to the outlet reference position 8. The inlet angle α.sub.2 is the angle between the tangent to the hub surface 20 at the reference position 1, and the axial direction.

(19) TABLE-US-00002 TABLE 2 Parameter Impeller Tip width (mm) 6.13 Diffuser Inlet width (mm) b2 5.4 Diffuser Length (mm) L 35.4 Diffuser Inlet angle α.sub.2 77.5

(20) Table 3 shows other parameters of the baseline configuration and the three embodiments, while Table 4 shows the axial width of the baseline configuration and the three embodiments at each of the radial positions 1 to 8.

(21) TABLE-US-00003 TABLE 3 Ratio of the distance from the rotational axis to throat Minimum Minimum position, to axial axial the distance extent of extent of of from the the gap the gap rotational Outlet at the as a % of Radial Normalised axis to the Outlet gap throat the position of radial radially- angle (mm) position impeller minimum position of outer edge Model (deg) α.sub.3 b3 44 (mm) tip width gap (mm) minimum gap of diffuser Baseline 90 4.318 4.318 70.4 62.45 116% 67.3% Embodiment 1 46.5 6.13 4.37 71.3 74.8 139% 80.6% Embodiment 2 77.5 6.13 4.88 79.6 70.1 130% 75.5% Embodiment 3 62 4.905 4.02 65.6 81.5 151% 87.7%

(22) TABLE-US-00004 TABLE 4 Baseline DOE2 DOE4 DOE13 Diff gap Diff gap Diff gap Diff gap Reference % of the % of the % of the % of the Reference Diff wheel Diff wheel wheel Diff wheel point gap tip width gap tip width Diff gap tip width gap tip width 1 5.42 88.418 5.42 88.418 5.42 88.418 5.42 88.418 2 4.32 70.473 4.75 77.488 4.97 81.077 4.72 76.998 3 4.32 70.473 4.48 73.083 4.88 79.608 4.36 71.126 4 4.32 70.473 4.38 71.452 4.91 80.098 4.14 67.537 5 4.32 70.473 4.42 72.104 5.01 81.729 4.04 65.905 6 4.32 70.473 4.6 75.041 5.2 84.829 4.04 65.905 7 4.32 70.473 4.97 81.077 5.49 89.560 4.18 68.189 8 4.32 70.473 5.97 97.390 6.08 99.184 4.83 78.793

(23) FIG. 4 shows the axial width of the baseline configuration and the three embodiments at each of the positions 1 to 8, according to table 4.

(24) FIG. 5(a) shows the relationship between the pressure ratio at the inlet and outlet, and the corporate mass flow for the base configuration and for embodiment 1. Corporate mass flow (shown as “corp mass flow” in FIGS. 5-7) is used here to mean the mass flow corrected for the inlet temperature and pressure. Line 101 shows the relationship for the baseline configuration, and a turbo speed of 65 k revolutions-per-minute (rpm). Line 102 shows the relationship for the baseline configuration, and a turbo speed of 95 k rpm. Line 111 shows the relationship for embodiment 1, and a turbo speed of 65 k rpm. Line 112 shows the relationship for embodiment 1, and a turbo speed of 95 k rpm. It can be seen that the pressure ratio is hardly different between embodiment 1 and the baseline configuration, except at the highest mass flows.

(25) FIG. 5(b) shows the efficiency as a function of corporate mass flow for the base configuration and for embodiment 1. Line 201 shows the relationship for the baseline configuration, and a turbo speed of 65 k rpm. Line 202 shows the relationship for the baseline configuration, and a turbo speed of 95 k rpm. Line 211 shows the relationship for the embodiment 1, and a turbo speed of 65 k rpm. Line 212 shows the relationship for embodiment 1, and a turbo speed of 95 k rpm. It can be seen that for low flow rates the baseline configuration and embodiment 1 have similar levels of efficiency. However, at the low turbo speed (65 k rpm), embodiment 1 is much more efficient than the baseline configuration for high flow rates. At the high turbo speed (95 rpm), embodiment 1 is slightly less efficient for high flow rates.

(26) FIG. 6(a) shows the relationship between the pressure ratio at the inlet and outlet, and the corporate mass flow for the base configuration and for embodiment 2. Line 101 shows the relationship for the baseline configuration, and a turbo speed of 65 k rpm. Line 102 shows the relationship for the baseline configuration, and a turbo speed of 95 k rpm. Line 121 shows the relationship for embodiment 2, and a speed of 65 k rpm. Line 122 shows the relationship for embodiment 2, and a turbo speed of 95 k rpm. It can be seen that the pressure ratio is hardly different between embodiment 2 and the baseline configuration, except at the highest mass flows.

(27) FIG. 6(b) shows the efficiency as a function of corporate mass flow for the base configuration and for embodiment 2. Line 201 shows the relationship for the baseline configuration, and a turbo speed of 65 k rpm. Line 202 shows the relationship for the baseline configuration, and a turbo speed of 95 k rpm. Line 221 shows the relationship for the embodiment 2, and a turbo speed of 65 k rpm. Line 222 shows the relationship for embodiment 2, and a turbo speed of 95 k rpm. It can be seen that for low flow rates the baseline configuration and embodiment 2 have similar levels of efficiency. However, at the low turbo speed (65 k rpm), embodiment 2 is much more efficient than the baseline configuration for high flow rates. At the high turbo speed (95 k rpm), embodiment 2 is slightly less efficient for high flow rates.

(28) FIG. 7(a) shows the relationship between the pressure ratio at the inlet and outlet, and the corporate mass flow for the base configuration and for embodiment 3. Line 101 shows the relationship for the baseline configuration, and a turbo speed of 65 k rpm. Line 102 shows the relationship for the baseline configuration, and a turbo speed of 95 k rpm. Line 131 shows the relationship for embodiment 3, and a turbo speed of 65 k rpm. Line 132 shows the relationship for the embodiment 3, and a turbo speed of 95 k rpm. It can be seen that the pressure ratio is hardly different between embodiment 3 and the baseline configuration.

(29) FIG. 7(b) shows the efficiency as a function of corporate mass flow for the base configuration and for embodiment 3. Line 201 shows the relationship for the baseline configuration, and a turbo speed of 65 k rpm. Line 202 shows the relationship for the baseline configuration, and a turbo speed of 95 k rpm. Line 231 shows the relationship for embodiment 3, and a turbo speed of 65 k rpm. Line 232 shows the relationship for embodiment 3, and a turbo speed of 95 k rpm. It can be seen that for low flow rates, and for high flow rates at the high turbo speed (95 k rpm), the baseline configuration and embodiment 3 have similar levels of efficiency. At the low turbo speed (65 k rpm), embodiment 3 is much more efficient than the baseline configuration for high flow rates.

(30) In summary, the embodiment 3 has efficiency improvement through the maps (though to a small extent at very high turbo speeds), whereas embodiments 1 and 2 only exhibit efficiency improvement at the low turbo speeds. On the other hand, for low turbo speeds, embodiments 1 and 2 show the greatest levels of efficiency improvement for high mass flow rates. All embodiments are more significantly more efficient than the baseline configuration at low turbo speed (about 65 k rpm) and high mass flow.

(31) Compared to the embodiments, the baseline configuration has a smaller diffusion length for flow mixing, but the diffusion process begins earlier (that is, at a radially inward position). The embodiments, by contrast, have an extended diffusion length for flow mixing, and the diffusion process is delayed. These factors produce better performance, especially at low speed.

(32) Although only a few embodiments of the diffuser have been described, many variations are possible within the scope of the invention as will be clear to a skilled reader.