US20150159660A1

US20150159660A1 - Axial turbine with radial vnt vanes

Info

Publication number: US20150159660A1
Application number: US14/099,793
Authority: US
Inventors: Vaclav Kares; Jeffrey Allen Lotterman; Josef Ferda
Original assignee: Honeywell International Inc
Current assignee: Honeywell International Inc
Priority date: 2013-12-06
Filing date: 2013-12-06
Publication date: 2015-06-11
Also published as: EP2881540A1

Abstract

A turbocharger including a turbine wheel having a hub-to-tip ratio of no more than 60% and blades with a high turning angle, a turbine housing forming an inwardly spiraling primary-scroll passageway that significantly converges to produce highly accelerated airflow into the turbine at high circumferential angles, and a two-sided parallel compressor. The compressor and turbine each produce substantially no axial force, allowing the use of minimal axial thrust bearings.

Description

The present invention relates generally to turbochargers and, more particularly, to an axial turbine having radial variable nozzle turbine vanes.

BACKGROUND OF THE INVENTION

With reference to FIG. 1, a typical turbocharger 101 having a radial turbine includes a turbocharger housing and a rotor configured to rotate within the turbocharger housing along an axis of rotor rotation 103 on thrust bearings and two sets of journal bearings (one for each respective rotor wheel), or alternatively, other similarly supportive bearings. The turbocharger housing includes a turbine housing 105, a compressor housing 107, and a bearing housing 109 (i.e., a center housing that contains the bearings) that connects the turbine housing to the compressor housing. The rotor includes a turbine wheel 111 located substantially within the turbine housing, a compressor wheel 113 located substantially within the compressor housing, and a shaft 115 extending along the axis of rotor rotation, through the bearing housing, to connect the turbine wheel to the compressor wheel.
The turbine housing 105 and turbine wheel 111 form a turbine configured to circumferentially receive a high-pressure and high-temperature exhaust gas stream 121 from an engine, e.g., from an exhaust manifold 123 of an internal combustion engine 125. The turbine wheel (and thus the rotor) is driven in rotation around the axis of rotor rotation 103 by the high-pressure and high-temperature exhaust gas stream, which becomes a lower-pressure and lower-temperature exhaust gas stream 127 and is axially released into an exhaust system (not shown).
The compressor housing 107 and compressor wheel 113 form a compressor stage. The compressor wheel, being driven in rotation by the exhaust-gas driven turbine wheel 111, is configured to compress axially received input air (e.g., ambient air 131, or already-pressurized air from a previous-stage in a multi-stage compressor) into a pressurized air stream 133 that is ejected circumferentially from the compressor. Due to the compression process, the pressurized air stream is characterized by an increased temperature over that of the input air.
Optionally, the pressurized air stream may be channeled through a convectively cooled charge air cooler 135 configured to dissipate heat from the pressurized air stream, increasing its density. The resulting cooled and pressurized output air stream 137 is channeled into an intake manifold 139 on the internal combustion engine, or alternatively, into a subsequent-stage, in-series compressor. The operation of the system is controlled by an ECU 151 (engine control unit) that connects to the remainder of the system via communication connections 153.
U.S. Pat. No. 4,850,820, dated Jul. 25, 1989, which is incorporated herein by reference for all purposes, discloses a turbocharger similar to that of FIG. 1, but which has an axial turbine. The axial turbine inherently has a lower moment of inertia, reducing the amount of energy required to accelerate the turbine. As can be seen in FIG. 2, the turbine has a scroll that circumferentially receives exhaust gas at the radius of the turbine blades and (with reference to FIG. 1) axially restricts the flow to transition it to axial flow. It thus impacts the leading edge of the turbine blades in a generally axial direction (with reference to col. 2).
For many turbine sizes of interest, axial turbines typically operate at higher mass flows and lower expansion ratios than comparable radial turbines. While conventional axial turbines generally offer a lower inertia, albeit with some loss of efficiency and performance, they suffer from an inability to be efficiently manufactured in the small sizes usable with many modern internal combustion engines. This is, e.g., due to the exceptionally tight tolerances that would be required, due to aerodynamic limitations, and/or due to dimensional limitations on creating small cast parts. Axial turbines also lack the ability to perform well at higher expansion ratios, such as are typically needed due to the pulsing nature of the exhaust of an internal combustion engine. Furthermore, conventional axial turbines have a significant change in static pressure across the blades, causing significant thrust loads on the thrust bearings of the rotor, and potentially causing blowby.
In some conventional turbochargers the turbines and compressors are configured to exert axial loads in opposite directions so as to lessen the average axial loads that must be carried by the bearings. Nevertheless, the axial loads from the turbines and compressors do not vary evenly with one another and may be at significantly different levels, so the thrust bearings must be designed for the largest load condition that may occur during turbocharger use. Bearings configured to support high axial loads waste more energy than comparable low-load bearings, and thus turbochargers that must support higher axial loads lose more energy to their bearings.
Variable nozzle turbines (“VNTs”) are known as preferable for some uses (such as with some diesel engines), For example, it is known to use a plurality of radial turbine vanes around the circumference of a radial turbine wheel, where the vanes rotate along various axial axes to open and close the radial inlet to the radial wheel. Likewise, it is known to use a plurality of axial turbine vanes axially in front of the inlet of an axial turbine wheel, where the vanes rotate along various radial axes to open and close the axial inlet to the axial wheel. Respectively, these examples would be considered a radial VNT and an axial VNT.
Accordingly, there has existed a need for a variable nozzle turbocharger turbine having a low moment of inertia, and characterized by a small size that does not require exceptionally tight tolerances, while having reasonable efficiency both at both lower and higher expansion ratios, and smaller axial loads. Preferred embodiments of the present invention satisfy these and other needs, and provide further related advantages.

SUMMARY OF THE INVENTION

In various embodiments, the present invention solves some or all of the needs mentioned above, typically providing a cost effective turbocharger turbine characterized by a low moment of inertia, and having a small size that does not require exceptionally tight tolerances, while operating at reasonable efficiency levels at both at both lower and higher expansion ratios, and having only small changes in static loads.
The invention provides a turbocharger configured to receive an exhaust gas stream from an engine configured to operate over a range of standard operating conditions, and to compress input air into a pressurized air stream. The turbocharger includes a turbocharger housing including a turbine housing, and a rotor configured to rotate within the turbocharger housing along an axis of rotor rotation. The rotor includes an axial turbine wheel, a compressor wheel, and a shaft extending along the axis of rotor rotation and connecting the turbine wheel to the compressor wheel. The turbine wheel is configured with a hub and a plurality of axial turbine blades configured to drive the rotor in rotation around the axis of rotor rotation when the turbocharger receives exhaust gas stream from the engine from a circumferential direction. The compressor wheel is configured to compress input air into the pressurized air stream.
Advantageously, the turbine housing forms an inwardly spiraling turbine primary-scroll passageway characterized by a significant enough radial reduction to accelerate exhaust gas such that a significant portion of the total pressure of the exhaust gas received by the turbine is converted into dynamic pressure. This allows an appropriately configured blade to extract a significant amount of energy from the exhaust gas without significantly changing the static pressure across the turbine blades. With a substantially unchanged static pressure across the turbine blades, the exhaust gas stream applies little to no axial pressure on the rotor.
The turbine wheel blades have an axially upstream edge, an axially downstream edge, a hub end, and a tip end opposite the hub end. The trailing edge is characterized by a radius at the hub end and a radius at the tip end. A feature of the invention is that the radius at the hub end of the turbine wheel trailing edge is no more than 60% of the radius of the tip end of the turbine wheel trailing edge. Further features include that the turbine wheel blades are limited to 16 or less in number, and are each characterized by a large turning angle.
Advantageously, these features provide for the extraction of a significant amount of energy from high-speed exhaust gas received in a highly circumferential direction without significantly impacting the static pressure of the gas. Furthermore, the turbine wheel does not require extremely tight manufacturing tolerances or small blade sizes, even when the wheel is manufactured in relatively small sizes.
The invention further features that, while the wheel is an axial turbine wheel, the turbine is provided with radial turbine vanes around the outer circumference of the inwardly spiraling turbine primary-scroll passageway. These vanes are configured to pivot around axial axes to open and close the outer circumference of the inwardly spiraling turbine primary-scroll passageway.
Advantageously, this configuration provides the advantages of the exercise of flow control through the turbine inlet, without having the vanes be active in the area where a significant portion of the total pressure of the exhaust gas received by the turbine has already been converted into dynamic pressure. This positioning of the vanes minimizes the effect of the vanes on the transition of the static pressure into dynamic pressure, particularly when the vanes are in a substantially or fully open configuration.
The invention further features that the compressor may be a two-sided, parallel, radial compressor including a compressor wheel with back-to-back oriented impeller blades including a first set of impeller blades facing axially away from the turbine and a second set of impeller blades facing axially toward the turbine. The compressor housing is configured to direct inlet air to each set of compressor blades in parallel. Advantageously, under this feature, the compressor is configured to produce substantially no axial load on the rotor. In combination with a turbine that also produces little or no axial load on the rotor, thrust bearing load levels can be significantly lower than in conventional turbochargers. The lower bearing load levels allow for the use of a more efficient thrust bearing, and thus increase the resulting overall efficiency of the turbocharger.
Other features and advantages of the invention will become apparent from the following detailed description of the preferred embodiments, taken with the accompanying drawings, which illustrate, by way of example, the principles of the invention. The detailed description of particular preferred embodiments, as set out below to enable one to build and use an embodiment of the invention, are not intended to limit the enumerated claims, but rather, they are intended to serve as particular examples of the claimed invention.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a system view of a prior art turbocharged internal combustion engine.

FIG. 2 is a cross-sectional front plan view of a turbocharger embodying the present invention.

FIG. 3 is a cross-sectional side view of the turbocharger depicted in FIG. 2, taken along line A-A of FIG. 2.

FIG. 4 is a plan view of certain critical flow locations relative to a turbine wheel depicted in FIG. 2.

FIG. 5 is a depiction of the camber of a turbine blade depicted in FIG. 2.

FIG. 6 is a perspective view of the turbine wheel depicted in FIG. 2.

FIG. 7 is a cross-sectional rear plan view of a portion of a turbine of the turbocharger depicted in FIG. 2.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

The invention summarized above and defined by the enumerated claims may be better understood by referring to the following detailed description, which should be read with the accompanying drawings. This detailed description of particular preferred embodiments of the invention, set out below to enable one to build and use particular implementations of the invention, is not intended to limit the enumerated claims, but rather, it is intended to provide particular examples of them.
Typical embodiments of the present invention reside in a motor vehicle equipped with a gasoline powered internal combustion engine (“ICE”) and a turbocharger. The turbocharger is equipped with a unique combination of features that may, in various embodiments, provide the aerodynamic benefits of a zero reaction turbine with the geometric benefits of a fifty percent reaction turbine, and/or provide significantly improved system efficiencies by combining less efficient components in a manner that reduces the bearing requirements, and thereby forms a system with a higher efficiency than the comparable unimproved system.
The turbine is configured to operate at reasonable efficiency levels at both lower and higher expansion ratios, having only small changes in static pressure across the turbine wheel (and thereby low rotor thrust loads), while it has a low moment of inertia, and is characterized by a small size, but does not require exceptionally tight tolerances. In combination with this, the compressor is also characterized by low axial thrust loads, providing for the turbocharger to require a thrust bearing that is significantly more efficient than is used in comparable conventional turbochargers.
With reference to FIGS. 2 & 3, in a first embodiment of the invention a typical internal combustion engine and ECU (and optionally an intercooler), such as are depicted in FIG. 1, are provided with a turbocharger 201 that includes a turbocharger housing and a rotor configured to rotate within the turbocharger housing along an axis of rotor rotation 203 on a set of bearings. The turbocharger housing includes a turbine housing 205, a compressor housing 207, and a bearing housing 209 (i.e., a center housing that contains radial and thrust bearings) that connects the turbine housing to the compressor housing. The rotor includes an axial turbine wheel 211 located substantially within the turbine housing, a radial compressor wheel 213 located substantially within the compressor housing, and a shaft 215 extending along the axis of rotor rotation, through the bearing housing, to connect the turbine wheel to the compressor wheel and provide for the turbine wheel to drive the compressor wheel in rotation around the axis of rotation.
The turbine housing 205 and turbine wheel 211 form a turbine configured to circumferentially receive a high-pressure and high-temperature exhaust gas stream from an exhaust manifold of the engine (such as the exhaust gas stream 121 from the exhaust gas manifold 123, as depicted in FIG. 1). The turbine wheel (and thus the rotor) is driven in rotation around the axis of rotor rotation 203 by the high-pressure and high-temperature exhaust gas stream acting on a plurality of blades 231 of the turbine wheel. The exhaust gas stream becomes a lower total pressure exhaust gas stream while passing through the blades, and is subsequently axially released via a turbine outlet 227 into an exhaust system (not shown).
The compressor housing 207 and compressor wheel 213 form a radial compressor. The compressor wheel, being driven in rotation by the exhaust-gas driven turbine wheel 211 (via the shaft 215), is configured to compress axially received input air (e.g., ambient air, or already-pressurized air from a previous-stage in a multi-stage compressor) into a pressurized air stream that may be ejected circumferentially from the compressor and sent on to an engine inlet (such as pressurized air stream 133 that is sent on to the engine inlet 139, as depicted in FIG. 1).
Turbine Volute
The turbine housing 205 forms an exhaust gas entrance passageway 217 leading into a primary-scroll passageway 219 configured to receive the exhaust gas stream from the engine in a direction normal to and radially offset from the rotor axis of rotation 203. The primary-scroll passageway forms a spiral adapted to significantly accelerate the speed of the gas stream to a high speed, which may be a supersonic speed for at least some operating conditions of the turbine (and its related engine). More particularly, the primary-scroll passageway turns the exhaust gas both inwardly around the axis of rotation 203 and axially toward the axial turbine wheel 211, thereby achieving (for some standard operating conditions of the engine) a supersonic flow having both a downstream axial component 221 and a downstream circumferential component 223.
Effectively, this configuration takes advantage of the conservation of angular momentum (rather than a convergent divergent nozzle) to achieve a high-speed airflow that may include a shockless transition to supersonic speeds for at least some operating conditions. Typically, a spiral characterized by a large radius change is required to achieve this change in velocity, and even though the resulting airstream is turned axially into an axial turbine wheel, it has a very high-speed circumferential component.
This circumferential component is achieved without the use of turning vanes, which would cause additional losses. Thus, the turbine inlet of this embodiment is of a vaneless design. As compared to a design with vanes, such a design advantageously is cost efficient, reliable (in that it eliminates parts from an environment in which they are likely to erode), avoids friction pressure losses, and avoids establishing a critical throat area that could choke the flow in some operating conditions.
With reference to FIGS. 2-4, this potentially supersonic flow of the accelerated exhaust gas stream in the inner radius of the primary-scroll passageway is directed into the turbine wheel 211. More particularly, the primary-scroll passageway is an inwardly spiraling passageway characterized by a primary-scroll inlet port 225 that connects the primary-scroll passageway to the exhaust gas entrance passageway 217. The primary-scroll passageway substantially forms a convergent passageway that spirals inward enough and converges enough to accelerate the exhaust gas, and to achieve supersonic speeds for at least some standard operating conditions of the engine (and thus of the turbocharger) as the exhaust gas turns axially downstream and impinges on the axially upstream end 233 of the blades 231.
The primary-scroll inlet port 225 is a planar location located along the passageways within the turbine that the exhaust gas travels through prior to reaching the turbine wheel. The location of the primary-scroll inlet port is defined relative to an opening in the passageway, which is characterized by a tongue-like shape when viewed in a cross-section taken normal to the rotor axis of rotation 203.
More particularly, the structure of a tongue 235 appears as a protrusion having a tip when viewed in the cross-section of FIG. 3. It should be noted that in some embodiments this structure will not vary in shape when the cross section is taken at different axial locations. In other embodiments the structure forming the tongue 235 may be shaped such that the location of the tip of the tongue varies when viewed in cross-sections taken at different axial locations.
The primary-scroll inlet port 225 is located at the tip of the tongue 235. To any extent that the circumferential location of the tip of the tongue appears to vary with the axial location of the cross-section considered, the primary-scroll inlet port 225 is defined to be at the most upstream location of the tip of the tongue, i.e., the upstream-most location at which the housing opens such that it is no longer radially interposed between the exhaust gas stream and the blades (even though the blades are axially offset from the exhaust gas stream). For the purposes of this application, the primary-scroll inlet port 225 is defined as the smallest planar opening from the exhaust gas entrance passageway 217 into the primary-scroll passageway 219, at the tip of the tongue. In other words, it is at the downstream end of the exhaust gas entrance passageway at the location at which the stream opens up to the blades.
The primary-scroll passageway 219 starts at the primary-scroll inlet port 225, and spirals inward 360 degrees around the axis of rotation to form a converging loop that rejoins flow coming in the primary-scroll inlet port 225. This convergent loop accelerates the exhaust gas circumferentially and turns it axially. Throughout the 360 degrees of the primary-scroll passageway 219, the accelerated and turned exhaust gas stream impinges on the blades 231, passing between the blades and driving the turbine wheel 211 in rotation.
In summary, the housing for the axial turbine wheel forms an inwardly spiraling primary-scroll passageway that surrounds the axis of rotor rotation. It begins at a primary-scroll inlet port 225 that is substantially radially external to the axially upstream ends of the blades, providing for the passageway to spiral inwardly and turn axially to accelerate the exhaust gas flow into the upstream ends of the axial turbine wheel blades.
Corrected Mass Flow
To provide for an adequate level of acceleration of the exhaust gas under the invention, the primary-scroll passageway 219 is configured with sizing parameters such that the corrected mass flow rate surface density of the turbine, when operated at a critical expansion ratio (E_cr), exceeds a critical configuration parameter, i.e., a critical corrected mass flow rate surface density (D_cr). More particularly, the sizing parameters for the scroll include a primary-scroll radius ratio (r_r) and a primary-scroll inlet port area (a_i), and are selected such that the corrected mass flow rate surface density of the turbine exceeds the critical configuration parameter D_crwhen the turbine is operated at the critical expansion ratio E_cr. These sizing parameters are defined relative to the primary-scroll inlet port 225, which is characterized by a centroid 237. For the gas to be axially adequately accelerated, this centroid will be substantially radially external to, and typically axially upstream of, an axially upstream end 233 of each blade 231.
The values of some of the above-recited terms are dependent upon the type of exhaust stream gas that will be driving the turbine. This exhaust-stream gas will be characterized by a Boltzmann Constant (k), and by a Gas Constant R-specific (R_sp). These constants vary by gas type, but for most gasoline powered engine exhaust gasses, the difference is anticipated to be small, with the constants being typically be on the order of k=1.3 and R_sp=290.8 J/kg/K.
The turbine housing has an ability to accelerate the exhaust gas that is characterized by the two sizing parameters recited above. The first sizing parameter, being the primary-scroll radius ratio r_ris defined to be a radius of a point 239 at the hub at leading edge of the turbine blades 231 (i.e., at the inner edge of the rotor inlet), divided by a radius of the centroid 237 of the planar area of the primary-scroll inlet port 225. The second, being the primary-scroll inlet port area a_iis defined to be the area of the primary-scroll inlet port 225.
As mentioned above, the geometry of this embodiment of a turbine is defined relative to operational parameters at the critical expansion ratio E_cr. This critical expansion ratio is obtained from the formula
$E_{cr} = {(\frac{k + 1}{2})}^{(\frac{k}{k - 1})}$
and is a function of the gas-specific Boltzmann's Constant k. A typical value for E_cris 1.832.
As recited above, the dimensions of the primary-scroll passageway 219 of this embodiment are limited by a primary-scroll radius ratio r_rand a primary-scroll inlet port area a_ithat cause the corrected mass flow rate surface density of the turbine to exceed the critical corrected mass flow rate surface density D_cr. This critical corrected mass flow rate surface density is obtained from the formula
$D_{cr} = r_{r} \frac{101325}{\sqrt{288 R_{sp}}} {(1 - \frac{(k - 1) {(r_{r})}^{2}}{(k + 1)})}^{(\frac{1}{k - 1})} \sqrt{\frac{2 k}{k + 1}}$
which varies with the primary-scroll radius ratio r_r.
For any given turbine, exactly one steady-state inlet condition for a given outlet static pressure (i.e., one inlet total pressure) will drive the turbine at a given expansion ratio such as the critical expansion ratio E_cr. A variation in the geometry of the volute, e.g., a variation of the radius ratio r_rand/or the primary-scroll inlet port area a_ican vary the steady-state mass flow rate that will drive the turbine at the given critical expansion ratio, and thus will affect the related corrected mass flow rate surface density.
If the primary-scroll radius ratio and the primary-scroll inlet port area are adequately selected, it will cause the corrected mass flow rate surface density at the primary-scroll inlet port 225 when driven at the critical expansion ratio E_crto be greater than the critical corrected mass flow rate surface density D_cr. While the relationships between the primary-scroll radius ratio, the primary-scroll inlet port area and the corrected mass flow rate surface density at the primary-scroll inlet port are complicated, and while they will typically be explored experimentally, it may be noted that in general, a higher radius ratio for the same port area will lead to a higher corrected mass flow rate surface density.
In an iterative method of designing a turbine under the invention, a person skilled in the art can first select a composition of an exhaust gas to be received from an engine, look up (from existing sources of gas properties) the related Boltzmann's Constant k and Gas Constant R_sp, and calculate the critical expansion ratio E_cr.
A first configuration of a turbine is then designed. The turbine includes a volute as described above, with an inwardly spiraling passageway that turns from a tangential direction to an axial direction, and an axial turbine wheel. The design is characterized by a first primary-scroll radius ratio r_r1and a first primary-scroll inlet port area a_i1.
A prototype is built, put on a gas stand, and run using the selected exhaust gas. The input total pressure is increased until a calculated expansion ratio reaches the critical expansion ratio E_cr. This expansion ratio is calculated from the total pressure at the inlet and the static pressure at the outlet. A steady state mass flow rate m, a total turbine inlet temperature T, and a total inlet pressure p_iare measured.
The corrected mass flow rate surface density is calculated from the measured data using the following formula:
$D_{ca} = \frac{m \times \sqrt{\frac{T}{288}}}{\frac{p_{i} \times a_{i}}{101325}}$
where a_iis the inlet port area. This calculated corrected mass flow rate surface density D_cais compared to the critical corrected mass flow rate surface density D_cr, which is calculated using the previously-identified formula. If the corrected mass flow rate surface density exceeds or equals the critical corrected mass flow rate surface density, then the design of an embodiment of the invention is complete. If the corrected mass flow rate surface density is less than the critical corrected mass flow rate surface density, then the design is considered insufficient to create the high-speed circumferential airflow needed under the invention, and another iteration of the design and testing steps are completed.
In this next iteration, the primary-scroll radius ratio r_rand/or the primary-scroll inlet port area a_iare appropriately adjusted (e.g., reduced) to increase the corrected mass flow rate surface density when taken at the critical expansion ratio E_cr. This process is repeated until a design is found in which the corrected mass flow rate surface density exceeds or equals the critical corrected mass flow rate surface density when taken at the critical expansion ratio E_cr.
In a potential alternative decision-making process for the above recited iterative design method, the decision to change one or both of the sizing parameters r_rand a_iis based on testing the axial loading of the turbine wheel (or the static pressure ratios that cause axial loading) by the exhaust gasses, over critical operating conditions (i.e., conditions that cause operation to occur at the critical expansion ratio E_cr). Another iteration is conducted if the axial forces are not below a threshold, such as the loading condition when the static pressure upstream of the wheel near the wheel hub is greater than 120% of the turbine outlet static pressure, i.e. the pressures differ at the most by 20% of the outlet pressure.
Wheel Blades
With reference to FIGS. 3-5, relative to the downstream axial flow component 221 and downstream circumferential flow component 223, each blade 231 is characterized by a lower surface 241 (i.e., the surface generally facing circumferentially into the downstream circumferential flow component) and an upper surface 243 (i.e., the surface generally facing circumferentially away from the downstream circumferential component).
The lower and upper surfaces of the blade 231 meet at a leading edge 245 (i.e., the upstream edge of the blade) and a trailing edge 247 (i.e., the downstream edge of the blade). The blades extend radially outward from a central hub 271 in a cantilevered configuration. They attach to the hub along a radially inner hub end 273 of the blade, and extend to a radially outer tip end 275 of the blade. The hub end of the blade extends from an inner, hub end of the leading edge to an inner, hub end of the trailing edge. The tip end of the blade extends from an outer, tip end of the leading edge to an outer, tip end of the trailing edge.
Typical axial turbines are typically provided with blades having blade lengths that are very small compared to the radius of the respective hub. Contrary to this typical convention, the present embodiment is provided with blades having a hub-to-tip ratio of less than or equal to 0.6 (i.e., the radius of the inner, hub end of the trailing edge is no more than 60% of the radius of the outer, tip end of the trailing edge).
While convention axial blades having high hub-to-tip ratios also require large numbers of blades to extract any significant amount of energy from the exhaust, the present blades are capable of extracting a very high percentage of the dynamic pressure of the high-speed highly tangential flow entering the turbine wheel. They can do so with a relatively limited number of blades, thereby limiting the rotational moment of inertia of the turbine wheel, and therefore providing for fast transient response time. Under numerous embodiments of the invention there are 20 or fewer blades, and for many of those embodiments there are 16 or fewer blades.
At any given radial location along the blade, the lower and upper surfaces are each characterized by a camber, and the blade is characterized by a median camber, which for the purposes of this application will be defined as a median camber curve 249 extending from the leading edge to the trailing edge at a median location equally between the upper and lower surfaces, wherein the median location is taken along a lines 251 extending from the upper camber to the lower camber, normal to the curve 249 along the median camber curve.
The median camber curve 249 comes to a first end at the leading edge 245. The direction of the median camber curve at the leading edge defines a leading-edge direction 253, and is characterized by a leading-edge direction angle β₁(i.e., a β₁blade angle) that is the angular offset between the leading-edge direction and a line that is parallel to the axis of rotation and passing through the leading edge (at the same radial location as the median camber), and therefore also to the downstream axial component 221 of the supersonic flow. The β₁blade angle is positive when the leading edge turns in to the circumferential flow component 223 (as depicted in FIG. 5), and zero when the leading edge faces directly along the axial flow component 221. The β₁blade angle can vary over the radial extent of the leading edge.
The median camber curve 249 comes to a second end at the trailing edge 247. The direction of the median camber curve at the trailing edge defines a trailing-edge direction 255, and is characterized by a trailing-edge direction angle β₂(i.e., a β₂blade angle) that is the angular offset between the trailing-edge direction and a line that is parallel to the axis of rotation and passing through the trailing edge (at the same radial location as the median camber). The β₂blade angle is positive when the trailing edge turns in to the circumferential flow component 223 (as depicted in FIG. 5), and zero when the trailing edge faces directly along the axial flow component 221. The blade angle β₂can vary over the radial extent of the trailing edge.
The sum of the β₁and β₂blade angles at a given radial location on a blade defines a turning angle for the blade at that radial location. The β₁+β₂turning angle can vary over the radial extent of the blade.
While the primary scroll efficiently accelerates the exhaust gas stream and thereby provides for a substantial increase in the dynamic pressure of the exhaust gas stream, it does not typically produce a flow with a high degree of axial uniformity, as might be seen from a vaned nozzle. The blades of the present embodiment, and particularly the shapes of their leading edges, are tailored so that each radial portion of the blade is best adapted to the flow that occurs at its radial location. This type of tailoring is not typical for conventional axial turbines, as they typically have vaned nozzles providing a high level of flow uniformity, and as they have a much higher hub-to-tip ratio that limits possible variations between the hub and tip flows.
Under the present embodiment, over the majority of the leading edge of each blade, the blade angle faces circumferentially upstream with respect to the axis of rotation (i.e., the β₁blade angle is positive). Moreover, the β₁blade angle is greater than or equal to 20 degrees (and possibly greater than or equal to 30 degrees) at both the hub end of the leading edge and the mid-span of the leading edge (i.e., the leading edge half way between its hub end and its shroud end). At the shroud end of the leading edge, the β₁blade angle is greater than or equal to −20 degrees (and possibly greater than or equal to −5 degrees).
Additionally, under the present embodiment, over the majority of the radial extent of each blade, the β₁+β₂turning angle is positive. Moreover, the turning angle is greater than or equal to 45 degrees at the hub end of each blade. The turning angle is greater than or equal to 80 degrees at the mid-span of each blade. At the shroud end of each blade, the β₁+β₂turning angle is greater than or equal to 45 degrees.
The chord line 261 (i.e., the line connecting the leading and trailing edge) has a positive angle of attack with respect to the downstream axial component 221, i.e., even though the leading-edge direction faces circumferentially upstream with respect to the axis of rotation, the chord line itself is angled circumferentially downstream with respect to the axis of rotation. In other words, the leading edge is circumferentially downstream of the trailing edge. This may vary in other embodiments.
The lower surface 241 of the blade of this embodiment is configured to be concave over substantially the full chord of the blade. Moreover, at the majority of radial locations, the lower surface is curved such that it has a range of locations 263 that are circumferentially downstream of both the leading edge and the trailing edge.
Static Pressure Drop
A key feature of the present embodiment of the invention is that it provides the inertial advantages of a typical axial turbine wheel (having a lower rotational moment of inertia than an equivalent radial turbine wheel), while it greatly enhances the ability of the axial turbine to extract the energy of the exhaust gas stream. To accomplish this, as previously suggested, the present embodiment is provided with a volute that uses conservation of angular momentum to efficiently accelerate the exhaust gas stream and convert a significant portion of the total pressure in the exhaust gas stream from static pressure to dynamic pressure, and further to provide the accelerated exhaust gas stream to an axial turbine wheel at a significant angle.
The turbine blade is configured to extract a significant portion of the energy of the dynamic pressure from the flow, but not to significantly change the static pressure of the flow. As a result of the volute converting a significant portion of the static pressure to dynamic pressure, and of the wheel extracting most of the dynamic pressure without changing the static pressure of the airstream, the turbine extracts a large percentage of the energy in the exhaust gas stream without receiving a significant axial load. A typical embodiment of the invention will be characterized by a static pressure change across the turbine wheel blades of less than ±20% of the static outlet turbine pressure across the turbine for at least some operating conditions of the range of standard operating conditions, thereby causing very little axial force to be applied to the turbine wheel. More particularly, the turbine is configured to limit the static pressure upstream of the wheel near the wheel hub to a value that is not greater than 120% of the turbine outlet static pressure, i.e. the pressures differ at the most by 20% of the outlet pressure. Some embodiments of the invention are characterized by substantially no static pressure drops across the rotor, thereby causing only a negligible axial force on the turbine wheel.
Wheel Hub
With reference to FIGS. 5 & 6, the radial size of the turbine wheel hub 271 varies along the blade inner hub end 273 from the leading edge 245 of each blade 231 to the trailing edge 247 of each blade, and it is uniform around the circumference. More particularly, the hub is radially larger at the leading edge than it is at the trailing edge, and the hub is radially larger at an intermediate axial location between the leading edge and the trailing edge than it is at either the leading edge or the trailing edge. This increase in thickness forms a smoothly continuous hump 277 that is axially close to the range of locations 263 on the blade lower surface 241 that are circumferentially downstream of both the leading edge and the trailing edge (i.e., where the median camber is parallel to the axial component of the flow).
The hump 277 is provided in a location in which significant diffusion occurs, and it prevents the diffusion from exceeding a critical level at which flow separation might occur. The potential for this problem is uniquely substantial because of the unique size and shape of the blades and the high level of kinetic energy of the flow. Because use of the hump helps avoid flow separation, the hump provides for improved efficiency over a similar wheel that lacks the hump.
Axially Balanced Compressor
With reference to FIG. 2, the compressor housing 207 and compressor wheel 213 form a dual, parallel, radial compressor. More particularly, the compressor wheel has back-to-back oriented impeller blades. A first set of impeller blades 301 are oriented in a conventional configuration with an inlet facing axially outward (away from the turbine) to receive air from that direction. A second set of impeller blades 303 are oriented in a reverse configuration with an inlet facing axially inward (toward the turbine) to receive air brought in tangentially and turned to travel axially into the second set of impeller blades. The first and second set of impeller blades can be manufactured in the form of a single, integral wheel, e.g., as illustrated, or may comprise an assembly of a plurality of parts.
The compressor housing 207 is configured to direct inlet air to each set of compressor blades in parallel, and to direct the passage of pressurized gas from each compressor. In this embodiment, the compressor housing comprises two separate axially positioned air inlets; namely, a first air inlet passage 305, that is positioned adjacent an end of the compressor housing to pass inlet air in an axial direction to the first compressor blades 301, and a second air inlet passage 307 that is separate from the first air inlet passage 305. Pressurized air that is provided by the compressor wheel 213 is directed radially from each set of impeller blades 301 and 303 through a single passage 311 to a compressor volute 313.
This dual-path, parallel, radial compressor configuration, while typically being less efficient than a comparable single-path radial compressor, will operate at higher speeds and might produce substantially no axial loading in steady state operation. The higher operating speeds will typically better match the operational speeds of the axial turbine.
Radial VNT Blades
Automotive turbochargers benefit from a different balance of size and inertia, as compared to performance, than many other turbine applications. More particularly, the small size and inertia may provide enhanced responsiveness, even if some performance is sacrificed. In some automotive applications such as diesel engines, it may be preferable to use a variable nozzle turbine (“a VNT”).
With reference to FIGS. 2, 3 and 7, the present embodiment of a turbocharger is provided with a plurality of turbine vanes 301. While the turbine wheel 211 is an axial turbine wheel, the turbine vanes are radial turbine vanes extending around the inwardly spiraling turbine primary-scroll passageway 219. These vanes are configured between an outer portion 511 of the passageway in which air primarily flows circumferentially and radially, and an inner, turning portion 513 of the passageway 219, in which the exhaust turns to flow in circumferentially and axially, and transitions to having a significantly lower static pressure. The vanes 501 pivot around axial axes 515 to open and close the connection between the outer portion 511 and the inner, turning portion 513 of the inwardly spiraling turbine primary-scroll passageway.
The vanes provide for the exercise of flow control through the turbine inlet, without the vanes extending through the flow immediately in front of the axial turbine blades, where a significant portion of the total pressure of the exhaust gas received by the turbine has already been converted into dynamic pressure. This positioning of the vanes minimizes the effect of the vanes on the transition of the static pressure into dynamic pressure, i.e., they allow the efficient use of conservation of angular momentum in generating a low static pressure swirl at the wheel inlet.
Variable radial vanes operate radially outside the circumference of the wheel. Unlike variable axial vanes, which have a vane clearance that varies with vane rotation, this configuration avoids the issue of vane clearance. Vane clearance variation can have negative effects on leakage losses. The vane clearance is generally the main loss contributor in the vane passage, and particularly when the vanes are in a near closed position where there can occur large leakage through the clearances. Radial vanes can maintain the same minimal clearance for all the vane positions while the axial vanes can have the minimal clearance only for closed position. If the axial vanes are opened a small amount then the clearance becomes bigger. A smaller number of vanes amplifies this effect, and increases leakage and thus losses.
It should be noted that the reuse of existing radial vane designs can significantly reduce the cost of developing radial vanes for the axial turbine wheels. In particular, known actuation mechanisms for radial turbine wheels can be adapted. Such an actuation mechanism is depicted in FIG. 7, but not in FIG. 2. An example of an actuation mechanism is disclosed in published U.S. Patent Application Number US20110171009, published Jul. 14, 2011 and filed Jan. 8, 2010, which is incorporated herein by reference for all purposes,
Synergies
The configuration of the present embodiment is significant for a number of reasons, and it is particularly effective for overcoming the efficiency limitations that limit the effectiveness of turbochargers on small gasoline powered engines, where the practical limitations of conventional axial turbines render them relatively ineffective for practical and efficient use.
The present invention provides an effective turbine with large blades that can be efficiently manufactured, even in small sizes. The comparatively large size and small number of axial turbine blades are well suited to casting in small sizes when smaller blades might be too small for conventional casting techniques. The high speed flow and large blades do not require manufacturing tolerances that may be limiting when applied to a very small turbine.
Singularly, the use of either a no-axial-load turbine or a no-axial load compressor is less efficient than their conventional axially loaded counterpart. Moreover, turbines and compressors are typically configured to have partially offsetting axial loads. Although these loads are far from perfectly matched, they do provide at least some relief from axial loads. If only one component (i.e., either the turbine or the compressor) creates no axial load, the remaining load from the other component is not partially offset, and even greater axial loads occur, requiring an even larger thrust bearing.
In the present invention, a no-axial-load compressor is combined with a no-axial-load turbine, allowing for the use of much more efficient thrust bearings. It is believed that in some embodiments the thrust load requirements may be as small as only 20% of the conventional counterparts. Bearings configured to carry such small loads can be adapted to be substantially more energy efficient. As a result, despite the potentially lower efficiencies of some of the system components, the overall system efficiency of the turbocharger may be significantly higher than in a conventional counterpart.
Finally, all these advantages can be adapted for turbines that typically benefit from having a variable nozzle through the unique combination of radial vanes with an axial turbine wheel.

OTHER ASPECTS

While many conventional turbochargers are designed to produce no downstream swirl, some embodiments of the present invention may be configured with blades that produce either negative or even positive swirl. In designing a turbine under the present invention, the production of downstream swirl might be considered of less interest than in the efficient extraction of energy while producing little or no axial loading.
It is to be understood that the invention comprises apparatus and methods for designing and producing the inserts, as well as for the turbines and turbochargers themselves. Additionally, the various embodiments of the invention can incorporate various combinations of the features described above. In short, the above disclosed features can be combined in a wide variety of configurations within the anticipated scope of the invention.
For example, while the above-described embodiment is configured as a forward-flow turbocharger (i.e., the exhaust gas stream is streamed through the turbine wheel so as to come axially out the end of the turbocharger), other embodiments may be configured with a reverse flow in which the exhaust gas stream passes through the turbine wheel in a direction toward the compressor. Such a configuration, while it might not fit in the standard spaces allotted for internal combustion engine turbochargers, exposes the bearing housing to less heat and pressure. Also, while the described embodiment uses a wheel with cantilevered (i.e., free-ended) blades that are radially surrounded by an unmoving housing shroud, other embodiments employing a shrouded wheel (i.e., a wheel having an integral shroud that surrounds the blades and rotates with them) is within the scope of the invention.
While particular forms of the invention have been illustrated and described, it will be apparent that various modifications can be made without departing from the spirit and scope of the invention. Thus, although the invention has been described in detail with reference only to the preferred embodiments, those having ordinary skill in the art will appreciate that various modifications can be made without departing from the scope of the invention. Accordingly, the invention is not intended to be limited by the above discussion, and is defined with reference to the following claims.

Claims

What is claimed is:

1. A turbocharger configured to receive an exhaust gas stream from an engine, and to compress input air into a pressurized air stream, comprising:

a housing including a turbine housing; and

a rotor configured to rotate within the housing along an axis of rotor rotation, the rotor including a turbine wheel, a compressor wheel, and a shaft extending along the axis of rotor rotation and connecting the turbine wheel to the compressor wheel, wherein the turbine wheel is an axial turbine wheel, and wherein the turbine housing forms an inwardly spiraling turbine primary-scroll passageway; and

a plurality of radial turbine vanes in the spiraling turbine primary-scroll passageway.