SIMPLIFIED VARIABLE GEOMETRY TURBOCHARGER WITH INCREASED FLOW RANGE

FIELD OF THE INVENTION

This invention addresses the need for a reduced cost, increased range, turbine flow control device, and accomplishes this by designing a simplified variable geometry turbocharger housing with controlled assymetric flow to the turbine wheel.

BACKGROUND OF THE INVENTION

Turbochargers are a type of forced induction system. They deliver air, at greater density than would be possible in the normally aspirated configuration, to the engine intake, allowing more fuel to be combusted, thus boosting the engine's horsepower without significantly increasing engine weight. A smaller turbocharged engine can replace a normally aspirated engine of a larger physical size, thus reducing the mass and aerodynamic frontal area of the vehicle.

Turbochargers (FIG. 1) use the exhaust flow (100) from the engine exhaust manifold, which enters the turbine housing at the turbine inlet (51) of a turbine housing (2), to drive a turbine wheel (70), which is located in the turbine housing. The turbine wheel is solidly affixed to one end of a shaft. A compressor wheel (20) is mounted to the other end of the shaft and held in position by the clamp load from a compressor nut. The primary function of the turbine wheel is providing rotational power to drive the compressor. Once the exhaust gas has passed through the turbine wheel (70) and the turbine wheel has extracted energy from the exhaust gas, the spent exhaust gas (101) exits the turbine housing (2) through the exducer (52) and is ducted to the vehicle downpipe and usually to after-treatment devices such as catalytic converters, particulate traps and NO_xtraps.

The power developed by the turbine stage is a function of the expansion ratio across the turbine stage, i.e., the expansion ratio from the turbine inlet (51) to the turbine exducer (52). The range of the turbine power is a function of, among other parameters, the mass flow through the turbine stage.

The compressor stage consists of a wheel and its housing. Filtered air is drawn axially into the inlet (11) of a compressor cover (10) by the rotation of the compressor wheel (20). The power generated by the turbine stage to the shaft and wheel drives the compressor wheel (20) to produce a combination of static pressure with some residual kinetic energy and heat. The pressurized gas exits the compressor cover (10) through the compressor discharge (12) and is delivered, usually via an intercooler, to the engine intake.

The design of the turbine stage is a compromise among: the power required to drive the compressor at different flow regimes in the engine operating envelope; the aerodynamic design of the stage; the inertia of the rotating assembly, of which the turbine is a large part, since the turbine wheel is manufactured typically in Inconel, which has a density 3 times that of the aluminum of the compressor wheel; the turbocharger operating cycle, which affects the structural and material aspects of the design; and the near field (exhaust flow) both upstream and downstream of the turbine wheel with respect to blade excitation.

Part of the physical design of the turbine housing is a volute (47), or pair of volutes, 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, ideally, 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 volute can be a perfect circle, defined as the base circle (71); or, in certain cases, such as a twin volute (48,49) as seen in FIG. 2A, it can describe a spiral of minimum diameter not less than 106% of the turbine wheel diameter.

The volute is defined by the decreasing radius of the outer boundary of the volute (53) and by the inner boundary, as described above, in one plane defined in the “X-Y” axis as depicted in FIG. 4, and the cross sectional areas, at each station, in the plane passing through the “Z” axis, as depicted in FIG. 8A. The “Z” axis is perpendicular to the plane defined by the “X-Y” axis and is also the axis of the turbine wheel.

Multiple entry volutes can also be created by dividing the volute area circumferentially. The volute is divided by axial walls (103, 104) which follow the decreasing outer boundary of the volute, as shown in FIG. 15A.

For consistency of product design, a system is used in which the development of the volute initiates at slice “A” (FIG. 4), which is defined as the datum for the remainder of the volute. The datum, slice “A”, is defined as the slice at an angle of “P” degrees above the “X-axis of the turbine housing containing the “X”-axis, “Y”-axis and “Z”-axis specifications of the volute shape.

The size and shape of the volute is defined in the following manner: The widely used term A/R represents the ratio of the partial area at slice “A” divided by the distance from the centroid (161) of the shaded flow area to the turbo centerline. In FIG. 8A, the position of the centroid (161) determines the distance R_Ato the turbo centerline. For different members of a family of turbine housings, the general shape remains the same, but the area at slice “A” is different, as is the distance R_A. The A/R ratio is generally used as the “name” for a specific turbine housing to differentiate that turbine housing from others in the same family (with different A/R ratios). In FIG. 8A, the volute is that of a typical divided turbine housing which forces the shapes of the volutes to be reasonably triangular and equal in area. In the case of a twin flow design (as depicted in FIG. 8A) the areas at slice “A” for both volutes are the same. The centroids (160, 161), of the areas are at the same radius R_A. The average centroid, (163), is on the turbine housing centerline at the same radius R_Asince the individual volutes at this section are symmetrical about the divider wall.

Slice “A” is offset by angle “P” from the “X”-axis. The turbine housing is then geometrically split into equal radial slices (often 30°, thus at (30x+P°)), and the areas (A_A-M) and the radii (R_A-M), along with other geometric definitions, such as corner radii are defined. From this definition, splines of points along the volute walls are generated, thus defining the full shape of the volute. The wall thickness is added to the internal volute shape, and, through this method, a turbine housing is defined.

The theoretically optimized volute shape for a given area is that of a circular cross-section since it has the minimum surface area which minimizes the fluid frictional losses. The volute, however, does not act on its own, but is part of a system; so the requirements of flow in the planes from slice “A”, shown in FIG. 4, to the plane at slice “M”, and from “M” to the tongue, influence the performance of the turbine stage. These requirements are often compromised due to other demands such as architectural requirements (space availability) outside the turbine housing, method of location and mounting of the turbine housing to the bearing housing, and the transition from slice “A” to the turbine foot (51), which combine to force turbine housing volutes to be of rectangular or triangular section, as well as in circular, or combinations of all shapes. The rectangular shape of the volute in FIG. 1, showing a section “D-K”, is a result of the requirement to not only to fit VTG (31) vanes into the space such that the flow is optimized through the vanes and that the vanes can be moved and controlled by devices external to the turbine housing, but also to minimize the outline of the turbine housing so the turbocharger fits on an engine.

The turbine housing foot is usually of a standard design as it mates to exhaust manifolds of many engines. The foot can be located at any angle to, or position relative to, the “volute”. The transition from the foot gas passages to the volute is executed in a manner which provides the best aerodynamic and mechanical compromise.

The roughly triangular shape of the volutes in FIG. 2, taken at the same sections as those above, is the more typical volute geometry for fixed and wastegated turbine housings. The addition of the divider wall (25) is to reduce aerodynamic “cross-flow” between the volutes in an effort to maintain pulse flow, from a divided manifold (36), to harvest the pulse energy in the work extracted by the turbine wheel. The pressure pulses in the exhaust manifold are a function of the firing order of the engine.

In commercial practice, turbine housings are typically designed in families (typically 5 to 7 in a family) which, in a given family, use turbine wheels of the same diameter, or a group of wheels with close to the same diameter. They may use the same turbine foot size, although this feature is sometimes customer driven. For example, a family of turbine housings for a 63mm turbine wheel may cover a range of A/Rs from 1.8 to 2.2. FIG. 5 depicts the area schedule for three volutes of a family. The largest volute is a 1.2 A/R volute, represented by the dotted line (45). The smallest volute is a 0.8 A/R volute; represented by the dashed line (46), and the mean volute, in the middle of the family, represented by the solid line. The X-axis depicts the angle of the slice from 30° (section “A”) to 360° (the tongue); the Y-axis depicts the area of the section at the respective angle. Typically there is an 8 to 10% difference in cross-sectional area (in the given case with 12 areas), at slice “A”, from one A/R to the next A/R in a design family. The volute outer wall of the largest A/R (45) of FIG. 5 is shown in FIG. 4 as the inner surface of the volute wall (40), and the smallest A/R (46) of FIG. 5 is shown in FIG. 4 as surface (41).

Some turbine wheels are specifically designed to harness this 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. This pulse energy is more predominant in commercial Diesel engines, which operate at around 2200 RPM with peak torque at 1200 to 1400 RPM, than in gasoline engines, which operate at much higher rotational speed, often up to 6000 RPM, with peak torque at 4000 RPM, such that the pulse is not as well defined.

The basic turbocharger configuration is that of a fixed turbine housing. In this configuration, the shape and volume of the turbine housing volute is determined at the design stage and cast in place. Most Diesel turbine housings are of the divided variety with a radial divider wall (25) as seen in FIG. 2 separating the two volutes to maintain the pulse energy to the turbine wheel. The divider wall length is typically such that the inner bound is approximately at the base circle. The closer the tip of the divider wall is to the base circle, the greater the preservation of pulse energy but the greater propensity for cracking of the casting in the divider wall. The reasons for this cracking are many but predominant are the dross which is pushed through the pattern at the casting process which means that the integrity of the material near the tip of the divider wall is less than optimal, and the second is the fact that the temperature distribution around the volutes causes the casting to want to “unwind”. The thermal forces generating the “unwinding” of the turbine housing are resisted by the vertical divider wall, the resultant being cracking in the wall. While a crack does little physical damage, the next step in cracking is for pieces of cast iron divider wall to separate from the casting and be ingested by the turbocharger or engine which can cause terminal damage.

The next level of sophistication, after that of the fixed turbine housing, is that of a wastegated turbine housing. In this configuration, the volute is cast in place, as in the fixed configuration above. In FIG. 2, the wastegated turbine housing features a port (54) which fluidly connects the turbine housing volute (49) to the turbine housing exducer (52). Since the port on the volute side is upstream of the turbine wheel (70), and the other side of the port, on the exducer side, is downstream of the turbine wheel, flow through the duct connecting these ports bypasses the turbine wheel (70), thus not contributing to the power delivered to the turbine wheel.

The wastegate in its most simple form is a valve (55) which can be a poppet valve, or a swing type valve similar to the valve in FIG. 2. Typically these valves are operated by a “dumb” actuator which senses boost pressure or vacuum to activate a diaphragm, connected to the valve, and operates without specific communication to the engine ECU. The function of the wastegate valve, in this manner, is to cut the top off the full load boost curve, thus limiting the boost level to the engine. This, in effect reduces the effective flow to the turbine, when desired (e.g. to prevent overdriving of the turbine), while allowing the full range of the turbine housing flow to the turbine wheel when full flow is desired. The wastegate configuration has no effect on the characteristics of the boost curve until the valve opens. More sophisticated wastegate valves may sense barometric pressure or have electronic over-ride or control, but they also have no effect on the boost curve until they actuate to open or close the valve.

FIG. 6A depicts the boost curve (65) for a fixed geometry turbine housing or a wastegated turbine housing in which the wastegate valve did not open. The X axis depicts mass flow; the Y axis depicts the pressure ratio. FIG. 6B depicts the boost curve (67) for a wastegated turbine housing of the same A/R as that for FIG. 6A in which the wastegate valve opened. In FIG. 6B, it can be seen that the lower shape (62) of the boost curve (67) is exactly the same as the boost curve (65) in FIG. 6A to the point (66) at which the valve opens. After this point, the boost curve is flat. While a wastegate can be used to limit boost levels, its turbine power control characteristics are rudimentary and coarse.

A beneficial byproduct of wastegated turbine housings is the opportunity to reduce the A/R of the turbine housings. Since the upper limit of the boost is controlled by the wastegate, a reduction in A/R can provide better transient response characteristics, while still controlling the upper limit. However, if the wastegated turbocharger has a “dumb” actuator, which operates on a pressure or vacuum signal only and is operated at altitude, then the critical pressure ratio at which the valve opens is detrimentally affected. Since the diaphragm in the actuator senses boost pressure on one side and barometric pressure on the other, the tendency is for the actuator to open later (since the barometric pressure at altitude is lower than that at sea level) resulting in over-boost of the engine. By introducing a smaller A/R turbine housing to take advantage of the wastegate, this A/R reduction also reduces the flow range of the turbine stage.

Engine boost requirements are the predominant drivers of compressor stage selection. The selection and design of the compressor is a compromise between: the boost pressure requirement of the engine; the mass flow required by the engine; the efficiency required by the application; the map width required by the engine and application; the altitude and duty cycle to which the engine is to be subjected; the cylinder pressure limits of the engine; etc.

The reason this is important to turbocharger operation is that the addition of a wastegate to the turbine stage allows matching to the low speed range with a smaller turbine wheel and housing. Thus, the addition of a wastegate brings with it the option for a reduction in inertia. Since a reduction in inertia of the rotating assembly typically results in a reduction of particulate matter (PM), wastegates have become common in on-highway vehicles. The problem is that most wastegates are somewhat binary in their operation, which does not fit well with the linear relationship between engine output and engine speed.

The next level of sophistication in boost control of turbochargers is the VTG (the general term for variable turbine geometry). Some of these turbochargers have rotating vanes and some have sliding sections or rings. Some titles for these devices include: variable turbine geometry (VTG); variable geometry turbine (VGT); variable nozzle turbine (VNT); or simply variable geometry (VG).

VTG turbochargers utilize adjustable guide vanes (FIGS. 3A and 3B) rotatably connected to a pair of vane rings and/or the nozzle wall. These vanes are adjusted to control the exhaust gas backpressure and the turbocharger speed by modulating the exhaust gas flow to the turbine wheel. In FIG. 3A the vanes (31) are in the minimum open position. In FIG. 3B the vanes (31) are in the maximum open position. The vanes can be rotatably driven by arms engaged in a unison ring, which can be located above the upper vane ring. For the sake of clarity, these details have been omitted from the drawings. VTG turbochargers have a large number of very expensive alloy components, which must be assembled and positioned in the turbine housing so that the guide vanes remain properly positioned with respect to the exhaust supply flow channel and the turbine wheel over the range of thermal operating conditions to which they are exposed. The temperature and corrosive conditions force the use of exotic alloys in all internal components. These materials are very expensive to procure, machine, and weld (where required). Since the VTG design can change turbocharger speed very quickly, extensive software and controls are a necessity to prevent unwanted speed excursions. This translates to expensive actuators. While VTGs of various types and configurations have been adopted widely to control both turbocharger boost levels and turbine backpressure levels, the costs of the hardware and implementation are high.

The cost of a typical VTG, in the same production volume, is from 270% to 300% the cost of the same size, fixed geometry turbocharger. This disparity is due to a number of pertinent factors from the number of components, the materials of the components, the accuracy required in the manufacture and machining of the components, to the speed, accuracy, and repeatability of the actuator. The chart in FIG. 7 shows the comparative cost for the range of turbochargers from fixed to VTGs. Column “A” represents the benchmark cost of a fixed turbocharger for a given application. Column “B” represents the cost of a wastegated turbocharger for the same application, and column “C” represents the cost of a conventional VTG for the same application.

Thus it can be seen that, for both technical reasons and cost drivers there needs to be a relatively low cost turbine flow control device which fits between wastegates and existing VTGs in terms of cost.

SUMMARY OF THE INVENTION

The present invention relates to a simplified, low cost, variable geometry turbocharger, and more particularly, a turbine flow controlling device, which uses a divided turbine housing with assymetric volute A/Rs coupled with a flow modulation device to change the effective exhaust mass flow to the turbine wheel, while increasing the turbine stage flow range. By controlling the mass flow of exhaust, which the turbine housing directs to the turbine wheel, with a set of asymmetrically configured volute cross sectional areas, and controlling the flow through the two volutes with a relatively simple flow controlling device, the flow range can be both broadened and controlled in a manner exceeding the range available with a symmetrically configured volute cross sectional areas without the flow controlling device.

The variable geometry turbocharger is simplified yet able to maintain pulse energy. In a first embodiment, a turbine housing is provided with a pivoting flow control valve which pivots about a point near the entry to the turbine housing. By moving the valve about the pivot point, the flow through the turbine housing is increasingly blocked at the large volute whereby flow is biased to the small volute and continues from there to the turbine wheel, thus causing the turbine housing, through effective loss of the larger volute, to act as a smaller A/R turbine housing. In the second embodiment of the invention, a rotating butterfly-configured flow control valve within the volute, which pivots about the center of the valve blade, pivots about said centerline to vary the flow from the large volute to the small volute and on to the turbine wheel, thus causing the turbine housing to act as a smaller A/R turbine housing.

Testing by the inventors determined that a 60/40 split of “A”-section areas with the hub side at 60% and the shroud side at 40% produced a desirable mass flow split with the restrictor valve fully open. The asymmetric turbine housing has a larger left or hub side volute (48) and a smaller or shroud side volute (49) situated axially about a divider wall (25).

BRIEF DESCRIPTION OF THE DRAWINGS

The present invention is illustrated by way of example and not limitation in the accompanying drawings in which like reference numbers indicate similar parts, and in which:

FIG. 1 depicts the section for a typical VTG turbocharger;

FIG. 2 depicts a pair of sections of a typical wastegated turbocharger;

FIGS. 3A,B depict a pair of sections of a typical VTG turbocharger;

FIG. 4 depicts a section of a typical fixed turbine housing showing construction radial lines;

FIG. 5 is a chart of cross-sectional area development;

FIGS. 6A,B depict the compressor maps for a typical fixed, and a wastegated turbocharger;

FIG. 7 is a chart showing turbocharger relative costs;

FIGS. 8A,B depict the sections of two volute types at slice “A”;

FIG. 9 depicts a view of an assymetric turbine housing on a manifold;

FIGS. 10A,B depict two section views of the restrictor device on a circumferentially divided housing;

FIGS. 11A,B depict two views of a variation of a restrictor device on a circumferentially divided housing;

FIG. 12 depicts a view of the cracking in a turbine housing;

FIG. 13 depicts a section view of closed slots in a turbine housing divider wall;

FIG. 14 depicts a section view of open slots in a turbine housing divider wall;

FIGS. 15A,B depict two views of the third embodiment on a radially divided housing;

FIG. 16 is a chart depicting mass flow;

FIG. 17 depicts the curtain areas for a sample of production turbine wheel diameters; and

FIG. 18 depicts the relationship between the crossflow area and D₃for different turbine stages.

DETAILED DESCRIPTION OF THE INVENTION

As discussed above, variable geometry mechanisms tend to double and more the cost of the basic turbocharger. The inventors sought the ability to modulate the exhaust flow to the turbine wheel in a more cost-effective manner. Therefore the inventors experimented with designs with divided volute areas, combined with a flow resistance device to provide both a cost and technically effective alternative for controlling the required wide range of exhaust gas flow to the turbine. In addition to the above gains, the inventors sought to provide a turbocharger matched to low flow regimes that would provide optimized turbo (and thus engine) transient response for low flow while still capable of delivering the high flows demanded by the engine in other than low flow conditions in the same, cost-effective turbocharger. This target keeps the gas velocities in the sweet spot which maximizes the stage efficiencies.

When a turbocharger is matched to the maximum flow requirement of an engine, the flow requirements across the entire engine operating regime are met. The problem is that matching the turbocharger to the maximum flow requirement means that the size of the turbine housing volute (and thus flow) is way too large for low engine flow regimes. The turbocharger's transient response characteristics are sluggish because the entire volute has to be filled in order to deliver flow to the turbine wheel. Since reducing the A/R of a turbocharger turbine housing to match the low flow requirement would mean that the turbocharger, operating within typical speed constraints, is not capable of providing sufficient flow for the high flow requirement of the upper end of the engine operating regimes, the inventors recognized the need to provide a novel variable geometry turbocharger. Furthermore with today's EGR (Exhaust Gas Recirculation) requirements, OEMs are running large amounts of EGR at part load (say 40% load) and no EGR at high speed yet they still desire, from a market standpoint, to deliver best-in-class power at full load. High EGR at low speed or part load requires low mass flow. Best-in-class power at the rated point with no EGR requires high mass flow so it can be seen that the turbine mass flow range needs to be capable of matching the flow requirements at these two extremes.

Turbine housing volute shapes and dimensions are defined by the area of section “A”, and all features and dimensions downstream of section “A” are controlled by the features and dimensions at section “A”. This system is used for consistency of design within the turbochargers designed and produced by a turbocharger manufacturer.

In accordance with the present invention, the inventors provide a novel turbine design able to produce a wider turbine flow range than would be available with volutes of equal area.

By controlling the mass flow through the turbine housing the inventors sought to control the mass flow of gas passing through the turbine housing to the turbine wheel. When the engine is operating in the low speed, low load condition, the boost level required to supply the required combustion gas (air) is relatively low. When the engine is in the high speed, high load condition, the boost level required to supply the engine under these load conditions is high. When the engine is transitioning from low load conditions to high load conditions, the turbocharger is required to supply an increasing volume of air at an increasing pressure ratio. Since the compressor stage is driven by the turbine stage, the mass flow of exhaust required to meet the engine (and thus the compressor) requirements has to change. That is, at the low load, low speed engine condition, the engine exhaust output, in terms of mass flow is low. At the high load, high engine speed condition, the engine output, in terms of mass flow is high. In the transition stage the exhaust mass flow has to change from low to high.

The problem is that the turbine stage must be matched to both of the above-described basic engine conditions, in addition to the requirements for EGR to allow the turbocharger to supply the requested flow and pressure ratio at any of these conditions. In order to force the turbocharger to change speed quickly, one experienced in the art would select a turbocharger with a small A/R turbine housing. In order to supply the required flow and pressure ratio at the high load, high speed condition one would select a turbocharger with a larger A/R turbine housing. The former small A/R turbine housing will provide good transient response characteristics, but insufficient mass flow to the turbine stage to generate the high speed, high load compressor requirement. The latter, large A/R turbine housing will provide the mass flow requirement to the turbine stage for the high speed, high load boost requirement but will not provide acceleration to the turbine wheel sufficiently quickly to produce acceptable transient response.

Obviously, it would be nice to have a system with two turbochargers, one larger and one smaller, and to be able to switch between the two. However, such a system would be expensive, would represent a large “heat sink”, would take much space in the engine compartment, and would add to the mass of the vehicle.

A properly matched small A/R turbine stage acting alone will provide acceptable transient response albeit at the expense of higher backpressure, compared to that of a turbine stage matched to the high load, high speed condition. In a non-EGR engine having high back pressure is a negative to the pressure differential across the engine and thus the efficiency of the engine. In a high pressure loop EGR engine configuration (as against a low pressure loop EGR engine configuration) the high back pressure in the exhaust system is part of the solution to drive the exhaust gas from the exhaust side of the engine into the inlet side of the engine which is seeing boost pressure. A large turbine housing A/R for a given set of engine parameters will develop lower exhaust back pressure than would a smaller A/R turbine housing under the same set of engine parameters. So being able to change the effective A/R of the turbine housing allows a single turbocharger to meet both the flow and back pressure requirements of a low speed, low load condition, and a high speed, high load condition.

By controlling the mass flow of exhaust, which the turbine housing directs to the turbine wheel, with a set of asymmetrically configured volute cross sectional areas, and controlling the flow through the two volutes with a relatively simple flow controlling device the flow range can be both broadened and controlled in a manner exceeding the range available with a symmetrically configured volute cross sectional areas without the flow controlling device.

After initially experimenting with a symmetrically divided volute turbine housing, the inventors next experimented with asymmetric divided turbine housings, and determined that by substituting one of the volutes with another volute of a smaller A/R that the flow range would drop and the maximum flow range through that volute would also drop. Similarly by replacing one volute with another volute of a larger A/R, the maximum flow range of that volute would rise. By putting together a combination of a larger and a smaller volute, and controlling the degree of blockage of the larger volute, the flow range of the inventive turbine housing exceeds that of the original prototype turbine housing with symmetrical divided volute. In FIG. 16 the bars (22) with the horizontal hatches represent the mass flow of a turbine housing with equal (50-50) area volutes and the bars (23) with the vertical hatches represent the mass flow of a turbine housing with assymetric (60-40) area volutes. While the mass flows with the restrictor valve fully open are equal to one another, when the restrictor valve is in the fully closed position (i.e., effectively blocking the flow into the larger volute) the mass flow of the assymetric A/R configuration is less than the mass flow of the equal A/R configuration. The sum of the areas at section “A” for both configurations is within 0% to 3%, while the change in mass flow is in the range of 10% to 13%.

FIG. 8A depicts a typical symmetrical turbine housing volute configuration in which the centroids (160, 161) of the two volutes are at the same radius, R_A, from the centerline. Since the turbine housing is symmetrical, the effective centroid (163) of both volutes lies in the divider wall between the volutes. FIG. 8B depicts an example in which the left volute is an A/R size larger than that of the symmetrical turbine housing of FIG. 8A, and the right volute is an A/R size smaller in area, thus 2 A/R sizes smaller than that of the left volute. In this case the centroid of the right volute is at a radius R_Cfrom the centerline and axially closer to the centerline of the turbine housing. The centroid of the left volute is at a radius R_Bfrom the centerline and axially further from the centerline of the turbine housing. The effective centroid (164) of both volutes in the turbine housing is now at a radius R_Doffset to the left of the divider wall centerline.

To produce an optimal asymmetric turbine housing the inventors looked at several options of volute sizes from one volute A/R up, or one volute down from equally sized volutes, to going from equal sized volutes and making the hub-side one volute A/R up and the shroud-side one volute A/R down. Testing by the inventors determined that the latter solution, which was a 60/40 split of “A”-section areas with the hub side at 60% and the shroud side at 40% produced the desired mass flow split with the restrictor valve fully open.

In all divided turbine housings there exists a cross-flow “curtain” between the tip of the divider wall, at its minimum diameter, and the tips of the turbine wheel. To minimize turbine wheel excitation caused by the action of the rotating turbine wheel blades passing the static tongue, (26) FIG. 4, at the start of the divider wall, as a rule of thumb, the depth of the divider wall typically does not extend closer to the turbine wheel blade tips than a ratio of from 120% to 150% of the turbine wheel tip diameter D₃. This ratio of D_bc/D3 is usually determined by company design rules and technical goals. The diameter D_bcis known as the base circle. Because turbine housing divider wall tips are prone to cracking, due to the dross of the molten cast iron being forced into the tip, the innermost, or minimum diameter of the divider wall is typically not less than 120% to 150% of the turbine wheel tip diameter. This “curtain” between the divider wall and turbine wheel allows for crossflow of exhaust gas, between the two volutes as well as cross-talk between the pulses in the exhaust flow, the latter being the reason for having the divider wall in the first place.

For a turbine stage with a base circle, the diameter of which is 120% of the turbine wheel diameter, there exists a cross flow “curtain”, with an area which is from 70% to 105% of the area of both volutes, in a symmetric configuration, at Section “A”, for a turbine housing family of 5 A/Rs. For a turbine stage with a base circle, the diameter of which is 150% of the turbine wheel diameter, the cross flow “curtain”, has an area which is from 199% to 299% of the area of both volutes, in a symmetric configuration, at Section “A”, for a turbine housing family of 5 A/Rs. From this analysis it can be seen that the curtain area can provide a very large cross sectional area for crossflow from one volute to the other.

Since the curtain area is a function of both the turbine wheel diameter D₃and the minimum position of the divider wall D_bc, the curtain area varies for different values of D₃. FIG. 17 depicts the curtain areas for a sample of production turbine wheel diameters from 64 mm to 96 mm. The curtain areas (133) are bracketed between or limited by an upper bound line (131) and a lower bound line (132). As would be expected the range of curtain areas increases as D₃increases. In FIG. 17 the turbine wheel diameters D₃are shown as (123) and a line (124) depicts the trend of turbine wheel diameters, D₃for the analyzed turbos. This chart also contains ratios of D_bc/D₃ranging from 1.25 to 1.35.

The inventors determined through testing with a 64 mm turbine wheel, that for an asymmetrically configured 60/40 volute combination with a restrictor valve, the optimum cross flow area, which includes ports in the divider wall plus the area of the base circle “curtain” (determined by the difference between the area under D_bcminus the area under D₃), was an area with a ratio of 289.6% that of a single, symmetrical, volute cross sectional area at Section “A” (i.e., half the area at Section “A”). This compares to the typical cross flow area of the same sized turbine housing with no slots or ports, with the same ratio of D_bc/D₃which has a crossflow area of only 182.6% of half the area at Section “A”.

As in the case of the relationship between the curtain areas and D₃and D_bc, the total cross flow area (122) is affected by not only D₃and D_bcbut also the variation in the area of a single volute at Section “A”. The crossflow areas (122) are bound by an upper bound line (126) and a lower bound line (127) The chart in FIG. 18 depicts the relationship between the crossflow area and D₃(123) for different turbine stages analyzed by the inventors.

To select a crossflow area, determine the value of D₃, the diameter of the turbine wheel in inches. The example is that of a 76 mm (2.992″) turbine wheel, shown as a horizontal line (128). From the turbine wheel diameter the vertical line (129) which intersects the turbine wheel diameter (123) cuts the lower bound line (127 and the upper bound line (128). The crossflow area is depicted as the vertical segment (130) of the vertical line (129) between the lower and upper bound lines (127 and 126).

The formula generating the data points, which are plotted on the charts shown in FIGS. 17 and 18, could be as follows:

$? = π / 4 (D_{bc}^{2} - D_{3}^{2})$

$\begin{matrix} Ratio = \frac{A_{CF}}{?} \\ = \frac{A_{curtain} + A_{slot}}{A 1 / 2_{Asect}} \\ = \frac{? + ?}{A 1 / 2_{Asect}} \end{matrix}$

$? indicates text missing or illegible when filed$

As depicted in FIG. 9, in the low flow condition, a pivotable valve member (72) is actuated to generate a flow restriction to the hub, or bearing housing side, larger volute (48) which forces the flow from the manifold through the shroud, or exducer side, smaller volute (49) to the turbine wheel (70).

The asymmetric turbine housing has a larger left or hub side volute (48) and a smaller right or shroud side volute (49) situated axially about a divider wall (25). A flow restrictor, in this case a pivotable valve member (72) is constrained within the joining faces of the manifold center section foot (37) and the turbine housing foot (51). While the inventors chose this configuration for cost and technical reasons the restrictor could be located in the hub side exhaust manifold passage (34).

As depicted in FIG. 10B, in the high flow condition the pivotable valve member (72) is in the central position, which biases neither the larger, hub side volute (48) nor the smaller, shroud side volute (49), to allow the maximum flow to the turbine wheel. In this high flow condition the pivotable valve member (72) of the flow restrictor device is aligned with the divider wall (25) of the turbine housing downstream the turbine foot (51). In the minimum flow position (depicted by the dashed line in FIG. 10B), the pivotable valve member (72) is rotated, preferably by a force exerted on the actuating arm (73), about the axis (74, 78) of the device, towards the closed position, such that it restricts the exhaust flow to the large volute (48) and causes the exhaust flow to flow through the small volute (49). The flow restrictor can be modulated to any position between fully open and fully closed.

A sectioned view of this version of the flow restrictor device is shown in FIG. 10A. In this view one can see that in the preferred embodiment of the invention, the flow restrictor blade is fabricated with two cylindrical bearing surfaces for pivoting and an actuating arm (73) for position control. One side of the cavity formed in the joint of the turbine housing foot (51) and exhaust manifold foot (37), which houses a bearing surface is a blind bore (77) while the other (75) is an open bore. On the open bore side a piston ring (76) provides not only an axial alignment for the flow restricting device, but also a gas seal. The placement of the actuating arm (73) can be optimized to meet architectural constraints.

The inventors realized that the ratio of boost-to-backpressure as well as the backpressure alone increased as a function of engine speed and load, at both sea level and at altitude, which made the flow restrictor device in the exhaust system an ideal controlling parameter. When the pivotable flow restrictor is rotated towards the closed position, the turbine housing acts as if it were a smaller A/R turbine housing than would exist with the flow restrictor in the open position. This causes the exhaust backpressure to rise which is necessary for EGR flow from the exhaust side of the engine to the inlet side of the engine. Thus the rotation of the flow restrictor can be used to develop a pressure differential (from the exhaust side of the engine, to the inlet side of the engine) to aid EGR flow from the exhaust side of the engine to the inlet side of the engine.

In the first embodiment of the invention, the effective mass flow to the turbine wheel is controlled by a flow restrictor which pivots about a point in the turbine housing inlet or foot such that in the open position the pivotable valve member (72) of the flow restrictor is in line with the divider wall (25) of the turbine housing minimizing the restriction to the exhaust flow. As more restriction, or less mass flow to the turbine wheel, is required the pivot arm (73) is actuated to rotate about its axis (74, 78) causing the pivotable valve member (72) to impede the flow of exhaust gas to the large volute (48), which causes a modulatable reduction in mass flow to the turbine wheel.

In a variation to the first embodiment of the invention, as depicted in FIGS. 11A and 11B, the flow restrictor takes the form of a butterfly valve (80) which reduces the moment on the pivot arm (81) enabling the potential use of a lower force, and thus lower cost actuator. In FIG. 11A, which is a section view of the first variation of the first embodiment of the invention the configuration of the bearing surfaces and piston ring are the same as in the first embodiment. In FIG. 11B, the pivot location, in the case of the butterfly configuration, is approximately in the center of the flow path to the large, or hub side volute (48) so that rotation of the butterfly (77), about its axis (74, 84) provides an adjustable flow restriction to the hub side volute (48) biasing the flow to the shroud side volute (49). In the case of this variation to the first embodiment of the invention, in the minimum flow restriction position, as depicted in FIG. 11B, the butterfly valve is aligned with the flow through the volute, such that the tips of the butterfly valve close or shadow the hub side volute (48). A butterfly valve solution has the advantage of low actuation loads, since the moments on the two sides of the pivot cancel each other.

When the flow restrictor is in the partially open position, flow from the shroud side (smaller) volute (49), to the hub side (larger) volute (48) can be further facilitated by either shortening the length of the divider wall (25), or by fabricating slots into the divider wall.

Typically, in the commercial Diesel world, where the product can be expected to run for a million miles, turbine housing divider walls are prone to cracking. The inventors realized an opportunity to mechanically minimize this propensity for cracking in the divider wall by introducing pre-cast stress-relieving features in the divider wall. FIG. 12B depicts a turbine housing viewed along section A-A of FIG. 12A. This sectioning is typically performed to evaluate the condition of the turbine housing following a thermal cycling qualification test in which the turbocharger is subjected to extreme temperature cycling in an effort to determine its resistance to cracking. In FIG. 12B the cracks (87) depicted are typical of a commercial Diesel type turbine housing in the divider wall area.

The inventors surmised that if “stress relievers” in the form of slots or ports were cast into the divider wall then these ports would not only minimize the propensity for cracking but also provide a flow path from the un-modulated shroud side volute to the modulated hub side volute under conditions of partial to full restrictor valve closure. This additional flow path provides flow to the turbine wheel over a greater circumferential distance or area than would be possible without the slots or ports.

In the second embodiment of the invention, as depicted in FIG. 13, the effective mass flow to the turbine wheel is controlled by a flow restrictor in an assymetric turbine housing with crossflow ports (88) fabricated in the divider wall. In the preferred second embodiment of the invention the area of said ports is bound by the leading edge radial (89), the trailing edge radial (90), the inner edge circular segment (92) and the outer edge spiral (91) for each port. The sum of the areas of the crossflow ports in the turbine housing is approximately equal to the area of the modulated volute at section “A”. What is important is the sum of the areas of the ports not the geometry of the ports.

In a variation to the second embodiment of the invention, as depicted in FIG. 14, the effective mass flow to the turbine wheel is controlled by a flow restrictor in an assymetric turbine housing with crossflow slots (95) fabricated in the divider wall. In the preferred second embodiment of the invention, the area of said slots is bound by the leading edge radial (98), the trailing edge radial (99), the outer edge spiral (91) for each port and the inner boundary by the base circle (71). The sum of the areas of the crossflow slots in the turbine housing is approximately equal to the area of the modulated volute at section “A”. What is important is the sum of the areas of the slots and the crossflow area inside the divider wall tip, not the geometry of the slots. In the preferred second embodiment of the invention the outer bound (97) of the slot can be characterized by a keyhole configuration as the outer termination of the slot to minimize the propensity of the slot to be a stress raiser and initiate cracking.

Multiple flow turbine housings with the volute divider wall parallel to the turbocharger axis, i.e., axial surfaces rather than radial surfaces as in the basic twin flow turbine housing are not uncommon. The inventors saw the opportunity to use similar logic for multiple flow turbine housings with assymetric volute areas accompanied by a flow restrictor to further cost effectively widen the flow range of a turbine stage with this type of turbine housing.

In the third embodiment of the invention, a triple flow turbine housing as depicted in FIG. 15A is preferably used. Two axial volute divider walls (103, 104) are fabricated into the turbine housing such that the ratio of flows through the unrestricted adjacent volutes, from outer to inner are approximately 70% to 20% to 10%. These proportions can be varied depending upon requirements. The ratio of flows is only important in that the sum of the open areas of the modulated volutes is equal to the area of the modulating restrictor valve. A flow restrictor valve is provided. The flow restrictor valve pivots about a point in the turbine housing inlet or foot such that in the open position, the blade (89) of the flow restrictor is flush with the turbine housing outer volute wall minimizing the restriction to the exhaust flow. As more restriction, or less mass flow, to the turbine wheel is required, the pivot arm (73) is actuated to rotate about its axis causing the blade (89) to impede the flow of exhaust gas to first the outer volute(s) followed by the center volute. In this manner the effective mass flow to the turbine wheel is controlled by a flow restrictor which enables a modulatable reduction in mass flow to the turbine wheel.

In a variation to the third embodiment of the invention, the dividing walls (106, 107) are slotted (108) to allow flow from the outer volutes to reach the inner volutes and then the turbine wheel (70). The slots (108) also allow for mass flow modulation but with a more consistent and favorable flow distribution to the turbine wheel.

Now that the invention has been described,

SIMPLIFIED VARIABLE GEOMETRY TURBOCHARGER WITH INCREASED FLOW RANGE

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