KR20140001912A - Simplified variable geometry turbocharger with increased flow range - Google Patents

Abstract

The variable geometry turbocharger is simple but can maintain pulse energy. In a first embodiment, the turbine housing has a rotary flow control valve that rotates about a point near the inlet of the turbine housing. By moving the valve around the pivot point, the effective volume of the turbine housing vortex chamber is varied, thus effectively reducing the exhaust gas volume of the vortex chamber and allowing the control of the exhaust gas flowing to the turbine wheel. In the second embodiment of the present invention, the rotating wedge portion in the vortex portion is rotated from the first position to the second position, and it is possible to control the exhaust gas flowing to the turbine wheel by changing the effective volume of the vortex chamber.

Description

Simplified variable geometry turbocharger with increased flow range {SIMPLIFIED VARIABLE GEOMETRY TURBOCHARGER WITH INCREASED FLOW RANGE}

The present invention achieves this need by reducing costs, addressing the need for an increased range of turbine flow control devices, and by constructing a simplified variable geometry turbocharger housing with controlled asymmetrical flow into the turbine wheel.

Turbochargers are a type of forced induction system. The turbocharger delivers air to the engine intake at a higher density than would be possible in a normal intake configuration to allow more fuel to burn, increasing engine horsepower without significantly increasing engine weight. Engines with small turbochargers can replace conventional intake engines of large physical size, thus reducing the mass and aerodynamic front area of the vehicle.

The turbocharger (FIG. 1) utilizes an exhaust flow 100 that flows from the engine exhaust manifold to the turbine housing at the turbine inlet 51 of the turbine housing 2, and thus the turbine wheel 70 located in the turbine housing. Drive. The turbine wheel is firmly attached to one end of the shaft. The compressor wheel 20 is mounted on the other end of the shaft and held in place by a fixed load from the compressor nut. The main function of the turbine wheel is to provide rotational force for driving the compressor. When the exhaust gas passes through the turbine wheel 70 and the turbine wheel obtains the energy extracted from the exhaust gas, the spent exhaust gas 101 exits the turbine housing 2 through the outlet 52 and exits the lower part of the vehicle. And is generally introduced into aftertreatment devices such as catalytic converters, particulate matter filtering devices and nitrogen oxide filtering devices.

The power generated through the turbine stage is a function of the expansion ratio across the turbine stage, that is, the expansion ratio from the turbine inlet 51 to the turbine outlet 52. The range of turbine power is a function of the mass flow through the turbine stage, among other parameters.

The compressor stage consists of a wheel and a housing of the wheel. The filtered air is axially introduced into the inlet 11 of the compressor cover 10 by the rotation of the compressor wheel 20. Power generated in the shaft and wheel through the turbine stage drives the compressor wheel 20 to produce a combination of static pressure and some residual kinetic energy and heat. The compressed gas exits the compressor cover 10 through the compressor outlet 12 and is usually delivered to the engine inlet via an intercooler.

The configuration of the turbine stage includes the power required to drive the compressor with different flow regimes at engine operating limits; Aerodynamic configuration of the stage; Inertia of the rotating assembly (the turbine is a large part of the rotating assembly since the turbine wheel is typically made of Inconel, which has a density three times the aluminum density of the compressor wheel); Turbocharger operating cycles affecting the structural and material aspects of the configuration; And the near filed (exhaust flow) of both the upstream and downstream of the turbine wheel for blade operation.

Part of the turbine housing's physical configuration is the vortices 47 or a pair of vortices, the function of which is the best of the inlet flow conditions, from the energy of the exhaust gas to the power generated through the turbine wheel. It is to control the inlet conditions of the turbine wheel to provide the most efficient power transmission combined with the transient response characteristics. In theory, the exhaust flow from the engine is transmitted in a uniform manner from the vortex to the vortex about the turbine wheel axis. To this end, ideally, the cross-sectional area of the vortex chamber is as perpendicular to the direction of flow as possible and is gradually and continuously reduced until the angle is zero. The inner boundary of the vortex chamber may be a complete circle defined by the base circle 71, or in some cases, as in the twin vortex chambers 48, 49, as shown in FIG. 2A, the inner boundary of the vortex chamber may be It may be described as a spiral in which the minimum diameter is at least 106% of the turbine wheel diameter.

The vortex chamber is, as described above, by the outer boundary and by the inner boundary of the vortex chamber 53 whose radius is reduced in one plane defined by the "XY" axis shown in FIG. 4, and by "Z" shown in FIG. 8A. "Defined by the cross-sectional areas at each position in the plane passing through the axis. The "Z" axis is perpendicular to the plane defined by the "X-Y" axis and is also the axis of the turbine wheel.

Multiple inlet vortex chambers may be created by dividing the vortex chamber area along the perimeter. The vortex chamber is divided by axial walls 103, 104 along the reduced outer boundary of the vortex chamber, as shown in FIG. 15A.

For consistency of the product configuration, a system is used in which the creation of the vortex chamber starts at slice “A” (FIG. 4), where section “A” is defined as a reference for the rest of the vortex chamber. Reference cross section "A" is defined as the cross section making an angle "P" degree over the "X" axis of the turbine housing including the "X" axis, "Y" axis and "Z" axis structures of the swirl chamber shape.

The size and shape of the vortex chamber is defined in the following manner: The widely used term A / R is the ratio of the partial area in section "A" divided by the distance from the center 161 of the darkened flow area to the turbo center line. Indicates. In FIG. 8A, the position of the center 161 defines the distance R _A to the turbo center line. The overall shape of the different turbine housings of the set of turbine housings is the same, but the area in cross section "A" differs as if the distance R _A is different. The A / R ratio is generally used as a “name” for a particular turbine housing to distinguish a particular turbine housing in the same set (with different A / R ratios) from other turbine housings. The vortex chamber of FIG. 8A is the vortex chamber of a typical split turbine housing in which the shapes of the vortex chamber are of suitable triangles and have the same area. In the twin flow configuration (as shown in FIG. 8A), the area in section “A” for the two vortex chambers is the same. The centers 160, 161 of the area are at the same radius R _A. Since the individual vortex chambers in this section are symmetrical about the dividing wall, the average center 163 is at the turbine housing centerline with the same radius R _A.

The section "A" is spaced apart from the "X" axis by an angle "P". The turbine housing is thus geometrically divided into the same radial cross sections (usually at 30 °, thus (30x + P °)), with areas A _{A -M} and radii R _AM equal to the corner radii. Defined with other geometric definitions. From these definitions splines of the points along the vortex chamber walls are created, thus defining the complete shape of the vortex compartment. The thickness of the wall is added to the internal vortex compartment shape and in this way the turbine housing is defined.

Since the circular cross section has a minimum surface area which minimizes fluid friction losses, the theoretically optimized vortex chamber shape at a given area is circular cross section. However, the vortex chamber does not work on its own, but is part of the system, so flow in the plane from the cross section "A" to the plane of the cross section "M" and from the cross section "M" to the tongue shown in FIG. The requirements affect the performance of the turbine stage. These requirements are due to other requirements, such as design requirements outside the turbine housing (space efficiency), how to position and mount the turbine housing in the bearing housing, and the connection from the cross section "A" to the turbine foot 51. Usually compromised, these requirements are combined such that the turbine housing vortex chambers have a circular cross section, or combinations of all shapes, as well as a rectangular or triangular cross section. The rectangular shape of the vortex chamber of FIG. 1 showing the section “DK” is for installing VGT 31 vanes in space so that flow through the vanes is optimized and the vanes can be moved and controlled by devices outside the turbine housing. In addition to the requirements, the result is a requirement for minimizing the contour of the turbine housing so that the turbocharger is installed on the engine.

The turbine housing foot is generally a standard configuration suitable for exhaust manifolds of many engines. This foot may be located at any angle or at any position relative to the “vortex chamber”. The transition from the foot gas passages to the vortex chamber is made in a manner that provides optimum aerodynamic and mechanical compromise.

The schematic triangular shape of the vortex chambers of FIG. 2, obtained in the same cross sections as the above cross sections, is a more typical vortex chamber shape for fixed wastegate turbine housings. The addition of the dividing wall 25 is intended to reduce the aerodynamic "cross flow" between the vortex chambers in order to maintain the pulse flow from the split manifold 36 so as to obtain pulse energy from the work obtained from the turbine wheel. will be. The pressure pulses on the exhaust manifold are a function of the engine ignition sequence.

In commercial practice, turbine housings typically consist of sets (typically 5 to 7 in one set), and in a given set, turbine housings use turbine wheels of the same diameter or a group of wheels close to the same diameter. . Turbine housings can sometimes use the same turbine foot size, depending on the customer. For example, a set of turbine housings with a 63 mm turbine wheel may include an A / R range of 1.8 to 2.2. 5 shows an area table for a set of three vortex chambers. The largest vortex chamber is the 1.2 A / R vortex chamber, indicated by dashed line 45. The smallest vortex chamber is the 0.8 A / R vortex chamber indicated by broken line 46, and the mean vortex chamber in the middle of the set is represented by the solid line. The X axis represents the cross section angle between 30 degrees (section “A”) and 360 degrees (tongue portion), and the Y axis represents the area of the cross section at each angle. Typically, one A / R and the next A / R in one configuration set have a difference of 8 to 10% in the cross-sectional area (in a given case with 12 areas) in cross section “A”. The vortex chamber outer wall of the largest A / R 45 of FIG. 5 is shown in FIG. 4 as the inner surface of the vortex chamber wall 40, and the smallest A / R 46 of FIG. 4 is shown.

Some turbine wheels are specially configured to use this pulsed energy and convert the pulsed energy to rotational speed. Thus, converting pressure and speed from the exhaust gas of the pulsed flow turbine wheel of the divided turbine housing is more useful than converting the pressure and speed from the steady state exhaust flow to the turbine wheel speed. This pulse energy is operated at approximately 2200 RPM with maximum torque at 1200 to 1400 RPM, rather than gasoline engines that operate at much higher rotational speeds, typically up to 6000 RPM with maximum torque at 4000 RPM, rather than gasoline engines where pulses are not very limited. More prominent in the engine.

The basic turbocharger configuration is that of a fixed turbine housing. In this configuration, the shape and volume of the turbine housing vortex chamber are determined at the design stage and at the casting site. Most diesel turbine housings are divided housings with radial splitting walls 25 separating the two vortex chambers to maintain pulse energy in the turbine wheel as shown in FIG. 2. The length of the dividing wall is generally the length such that the inner boundary is approximately the base circle. The closer the stage of the dividing wall is to the base circle, the more the conservation of pulse energy is increased, but the cracking tendency of the casting of the dividing wall is increased. The causes of these cracks vary, but the main cause is the residue injected through the pattern of the casting process (which means that the integrity of the material adjacent to the end of the dividing wall is not optimal), and the second is the temperature around the vortex chambers. It is a fact that the distribution causes "unwind" of the casting. The thermal force that causes the "loosening" of the turbine housing is resisted by the vertical dividing wall, resulting in a crack in the wall. The cracks do some physical damage, but the next stage of the cracks may cause pieces of the cast iron splitting wall to separate from the casting and be sucked into the turbocharger or engine, resulting in fatal damage.

The next step after upgrading the stationary turbine housing is to upgrade the wastegate turbine housing. In this configuration, the vortex chamber is cast in situ as in the fixed configuration described above. The wastegate turbine housing in FIG. 2 features a port 54 that fluidly connects the turbine housing vortex chamber 49 to the turbine housing outlet 52. Since the port on the vortex chamber side is upstream of the turbine wheel 70 and the other side of the port on the outlet side is downstream of the turbine wheel, the flow through the duct connecting these ports bypasses the turbine wheel 70, Thus, it does not contribute to the power delivered to the turbine wheels.

The simplest type of wastegate is valve 55, which may be a poppet valve or a swing type valve similar to the valve of FIG. Typically, these valves sense a propulsion pressure or vacuum that actuates the diaphragm, are connected to the valve, and are also operated by a "dumb" actuator that operates without special communication with the engine electronic control unit (ECU). . In this way, the function of the wastegate valve cuts off the top of the full load boost curve, thus limiting the engine's propulsion level. This allows for full flow of the turbine housing to the turbine wheel when full flow is needed, while actually reducing the effective flow to the turbine when needed (eg to prevent excessive drive of the turbine). The wastegate configuration does not affect the properties of the propulsion curve until the valve is opened. More sophisticated wastegate valves can measure air pressure, or be equipped with an electronic over-ride or electronic control device, but these devices also do not affect the propulsion curve until the valve is operated to open or close. Do not.

FIG. 6A shows the propulsion curve 65 for a fixed geometry turbine housing or a wastegate turbine housing in which the wastegate valve is not open. The X axis represents mass flow and the Y axis represents pressure ratio. FIG. 6B shows the propulsion curve 67 for the wastegate turbine housing of the same A / R as in FIG. 6A with the wastegate valve open. In FIG. 6B, it can be seen that the lower shape 62 of the propulsion curve 67 is exactly the same as the propulsion curve 65 of FIG. 6A to the point 66 at which the valve opens. After this point, the propulsion curve is flat. The wastegate can be used to limit the propulsion level, but the turbine power control characteristics of the wastegate are the most basic and coarse.

A beneficial byproduct of wastegate turbine housings is the opportunity to reduce the A / R of the turbine housings. Since the upper limit of propulsion is controlled by the wastegate, the reduction in A / R can provide good transient response characteristics while controlling the upper limit. However, if the wastegate turbocharger is equipped with "dumb" actuators that operate only on pressure or vacuum signals and operate on high ground, the critical pressure ratio at which the valve opens is adversely affected. Since the diaphragm of the actuator measures the propulsion pressure on one side and the air pressure on the other side, the actuator tends to open later (since the air pressure at the altitude is lower than sea level), which in turn causes abnormal propulsion 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 propulsion requirements are the main factors in choosing a compressor stage. The choice and configuration of the compressor includes: the propulsion pressure requirements of the engine; Mass flow required for the engine; Efficiency required for application; Layout width required for the engine and application; Altitude and utilization at which the engine is applied; The compromise is made between the cylinder pressure limit of the engine and the like.

This is important for turbocharger operation because adding wastegates to the turbine stage makes it possible to cope with low speed ranges with smaller turbine wheels and housings. Thus, the addition of the wastegate brings the option to reduce the inertia. As the inertia reduction of the rotating assembly generally results in a reduction of particulate matter (PM), wastegates have become common in high speed vehicles. The problem is that most wastegates are somewhat dual in their operation, which does not fit well with the linear relationship between engine power and engine speed.

The next step in advancing the turbocharger's propulsion control is VTG (general term variable turbine geometry). Some of these turbochargers have rotary vanes, and some have sliding faces or rings. Some names for these devices are: Variable Turbine Geometry (VTG); Variable geometry turbine (VGT); Variable nozzle turbine (VNT); Or simply include variable geometry (VG).

The VTG turbocharger utilizes a pair of vane rings and / or adjustable guide vanes (FIGS. 3A and 3B) rotatably connected to the nozzle wall. These vanes are adjusted to control exhaust back pressure and turbocharger speed by varying the exhaust flow to the turbine wheel. In FIG. 3A, the wings 31 are in the minimum open position. In FIG. 3B, the wings 31 are in the maximum open position. The wings can be rotatably driven through arms fastened to a Unison ring that can be positioned above the upper wing ring. These details are omitted in the drawings for clarity. VTG turbochargers are equipped with a large number of expensive alloy parts, which keep the guide vanes properly positioned relative to the exhaust gas supply flow passage and turbine wheel in the range of thermal operating environments in which the parts are exposed. To be assembled and positioned in the turbine housing. Temperature and corrosive environments lead to the use of a special alloy on all internal components. These materials are expensive to buy, process and weld (if necessary). Since the VTG design can change turbocharger speed quickly, a lot of software and controls are needed to prevent unwanted speed deviations. This leads to expensive actuators. Although various types and configurations of VTGs are widely employed to control both turbocharger propulsion levels and turbine back pressure levels, hardware and implementation costs are high.

The cost of a typical VTG is from 270 to 300% of the cost of a fixed geometry turbocharger of the same size, at the same yield. This difference is due to a number of factors related to the number of parts, the material of the parts, the precision required for the manufacture and processing of the parts, to the speed, accuracy and repeatability of the actuator. The chart of FIG. 7 shows the relative costs of turbochargers ranging from fixed turbochargers to VTG. Pillar "A" represents the reference cost of a fixed turbocharger in a given application. Pillar "B" represents the cost of a wastegate turbocharger in the same application, and pillar "C" represents the cost of a typical VTG in the same application.

Therefore, both technical and cost factors require that a relatively inexpensive turbine flow control device be fitted between the wastegate and the existing VTG in terms of cost.

The present invention relates to a simple and low cost variable geometry turbocharger, and more particularly to an asymmetric vortex chamber combined with a flow control device to increase the turbine stage flow range while varying the substantial exhaust mass flow to the turbine wheel. A turbine flow control device employing a split turbine housing with A / Rs. By controlling the mass flow of the exhaust of the turbine housing directed to the turbine wheel through a set of asymmetrically configured vortex chamber cross sections, and by controlling the flow through the two vortex chambers through a relatively simple flow control device, the flow range is controlled by flow control. Symmetrically configured vortex chamber cross sections without device can be expanded and controlled in ways exceeding the possible range.

The variable geometry turbocharger is simple but can maintain pulse energy. In the first embodiment, the turbine housing has a pivotal flow control valve that rotates about a point near the inlet of the turbine housing. By moving the valve around the pivot point, the flow through the turbine housing is increasingly blocked in the large vortex chamber, biasing the flow to the small vortex chamber, leading to the turbine wheel in the small vortex chamber, and thus through the effective loss of the large vortex chamber. The turbine housing serves as a smaller A / R turbine housing. In a second embodiment of the present invention, a rotary butterfly-shaped flow control valve in a vortex chamber that rotates about the center of the valve blade rotates about the centerline to change the flow from the large vortex chamber through the small vortex chamber to the turbine wheel. Thus allowing the turbine housing to act as a small A / R turbine housing.

The inventors found that the 60/40 division of the "A" cross-sectional areas with 60% center side and 40% outside has obtained the desired split of the mass flow with the restrictor valve fully open. The asymmetric turbine housing has a larger left or center side vortex chamber 48 and a smaller or outer vortex chamber 49 positioned axially with respect to the dividing wall 25.

The invention is shown by way of example and not by way of limitation in the figures of the drawings in which like reference numerals designate like parts.
1 shows a cross section of a typical VTG turbocharger,
2 shows a pair of cross sections of a typical wastegate turbocharger,
3A and 3B show a pair of cross sections of a typical VTG turbocharger,
4 shows a cross section of a typical fixed turbine housing showing a structural radius,
5 is a chart of the improved cross-sectional area,
6A and 6B show a compressor map of a typical fixed turbocharger and wastegate turbocharger,
7 is a chart showing the relative costs of a turbocharger,
8A and 8B show cross sections of two vortex chamber types in cross section “A”,
9 shows an asymmetric turbine housing on a manifold,
10a and 10b show two cross sections of the limiting device on the housing divided in the circumferential direction;
11a and 11b show two variants of the limiting device on the housing divided in the circumferential direction,
12 shows a crack in the turbine housing,
13 shows a cross section of the closed slots of the turbine housing dividing wall,
14 shows a cross section of the open slots of the turbine housing splitting wall,
15a and 15b show two third embodiments on a radially divided housing,
16 is a chart showing mass flow;
17 shows the membrane areas for a sample of produced turbine wheel diameters,
18 shows the relationship between the cross flow area and D ₃ of the various turbine stages.

As mentioned above, variable geometry mechanisms tend to cost more than twice as much as a basic turbocharger. The inventors have sought to be able to regulate the exhaust flow to the turbine wheels in a more cost effective manner. Accordingly, the inventors have experimented with configurations with divided vortex compartment areas combined with flow resistance devices to provide both a cost and technically effective alternative to control the wide range of exhaust gas flow required to the turbine. In addition to the above improvements, the inventors can provide an optimized turbo (and thus engine) transient response to low flow while also adapting a turbo to a low flow scheme that can also deliver the high flows required by the engine in addition to the low flow conditions of the engine. We tried to provide a charger, a cost-effective turbocharger. This goal maintains the gas velocity at the optimum state to maximize stage efficiency.

If the turbocharger meets the engine's maximum flow requirements, flow requirements are met across the entire engine operating system. The problem is that fitting the turbocharger to the maximum flow demand means that the size of the turbine housing vortex chamber (and thus flow) is too large in low engine flow regimes. The transient response characteristics of turbochargers are sluggish because the entire vortex chamber must be filled to deliver flow to the turbine wheel. Reducing the A / R of the turbocharger turbine housing to meet low flow demands means that turbochargers operating within typical speed limits cannot provide sufficient flow for high flow demands on top of the engine operating system. The inventors have recognized the need to provide a new variable geometry turbocharger. Also, due to current EGR (exhaust gas recirculation) requirements, OEMs operate with high EGR at partial load (e.g. 40% load) and without EGR at high speed, but OEMs are still from the market perspective. You want to deliver the best power at full load. High EGR at low speed or partial load requires low mass flow. It can be seen that the highest power at the rated point without EGR requires high mass flow, and therefore the turbine mass flow range must be able to meet the flow conditions at these two extremes.

Turbine housing vortex chamber shapes and dimensions are defined by the cross-sectional area of cross section "A", and all features and dimensions downstream of cross section "A" are adjusted by the features and dimensions in cross section "A". This system is used for consistency of the configuration of turbochargers constructed and produced by turbocharger manufacturers.

In accordance with the present invention, the inventors provide a new turbine configuration that can create a wider turbine flow range than can be used through vortex chambers of the same area.

By controlling the mass flow through the turbine housing, the inventors wanted to control the mass flow of gas entering the turbine wheel through the turbine housing. When the engine is operated at low speed, low load conditions, the propulsion level required to supply the required combustion gases (air) is relatively low. When the engine is operated at high speed, high load conditions, the propulsion level required to supply the engine under these load conditions is high. When the engine is switched from low load conditions to high load conditions, the turbocharger is required to supply an increased amount of air at an increased pressure ratio. Since the compressor stage is driven by a turbine stage, the mass flow of the exhaust required to meet engine (and thus compressor) requirements must vary. That is, at low load, low speed engine conditions, the engine exhaust power is low in terms of mass flow. At high load, high speed engine conditions, engine power is high in terms of mass flow. In the conversion stage, the exhaust mass flow must change from low to high flow.

The problem is that the turbine stage must meet all of the basic engine conditions described above, in addition to the requirements for the EGR to allow the turbocharger to supply the required flow and pressure ratios in all these conditions. In order for the turbocharger to change speed rapidly, one skilled in the art can select a turbocharger with a small A / R turbine housing. To supply the required flow and pressure ratios under high load, high speed conditions, one skilled in the art can choose a turbocharger with a large A / R turbine housing. The small A / R turbine housing, the former, provides good transient response characteristics but will provide insufficient mass flow to the turbine stage to generate high speed, high load compressor requirements. The latter large A / R turbine housing provides mass flow requirements to the turbine stage for high speed, high load propulsion requirements, but will not accelerate the turbine wheels fast enough to produce an acceptable transient response.

Obviously, it is desirable to have a system with two turbochargers, one larger and one smaller, which can be switched between the two. However, the system is expensive, exhibits a large "heat sink", takes up a lot of space in the engine room, and will add to the mass of the vehicle.

Properly matched small A / R turbine stages that operate alone will provide acceptable transient response even at the expense of higher back pressure than turbine stages for high load, high speed conditions. In engines without EGR, having a high back pressure is negative for the pressure differential of the engine and therefore poor for engine efficiency. In the high pressure loop EGR engine configuration (as opposed to the low pressure loop EGR engine configuration), the high back pressure of the exhaust system is part of the solution for supplying the exhaust gas from the exhaust side of the engine to the intake side of the engine under propulsion pressure. Large turbine housing A / R for a given set of engine parameters will produce lower exhaust back pressure than smaller A / R turbine housing under the same set of engine parameters. Thus, being able to change the effective A / R of the turbine housing allows one turbocharger to meet both low speed, low load conditions and high speed, high load conditions of flow and back pressure requirements.

By controlling the mass flow of the exhaust from the turbine housing to the turbine wheel through a set of asymmetrically configured vortex chamber cross sections and controlling the flow through the two vortex chambers through a relatively simple flow control device, the flow range is controlled by the flow control device. Symmetrically configured vortex compartment cross sections can be expanded and controlled in ways beyond the possible range.

After the first experiments with symmetrically divided vortex turbine turbine housings, the inventors did the following experiments with asymmetrically divided turbine housings and replaced the flow range by replacing one of the vortex chambers with another smaller vortex chamber. Was found to be reduced and the maximum flow range through this vortex could also be reduced. Likewise, by replacing one vortex chamber with another vortex chamber of larger A / R, the maximum flow range of the vortex chamber can be increased. By combining the large vortex chamber and the small vortex chamber together and controlling the degree of interruption of the large vortex chamber, the flow range of the turbine housing of the present invention exceeds the flow range of the original circular turbine housing with symmetrically divided vortex chambers. . In FIG. 16, the bars 22 formed with horizontal hatching lines represent the mass flow of the turbine housing with vortex chambers of the same (50-50) area, and the bars 23 formed with vertical hatching lines are asymmetric (60). -40) represents the mass flow of a turbine housing with vortex chambers of area. The mass flows with the fully open restriction valves are identical to each other, while if the restriction valves are in the fully closed position (i.e. effectively blocking the flow to the large vortex chamber), the mass flows in the asymmetrical A / R configuration will be identical. Smaller than the mass flow of the composition. The sum of the cross-sectional areas of the section “A” for both configurations is within 0 to 3%, but the change in mass flow is in the range of 10 to 13%.

8A shows a typical symmetric turbine housing vortex compartment configuration in which the centers 160, 161 of the two vortex chambers are at the same radius R _A from the center line. Since the turbine housing is symmetrical, the effective center 163 of the two vortex chambers lies in the dividing wall between the vortex chambers. FIG. 8B shows an A / R size larger than the A / R size of the left vortex chamber of the symmetric turbine housing of FIG. 8A and the right vortex chamber is an A / R size smaller in area. An example in which the R size is smaller than the A / R size of the left vortex chamber is shown. In this case, the center of the right vortex chamber is at a radius R _C from the center line and is closer in the axial direction to the center line of the turbine housing. The center of the left vortex chamber is at the radius R _B from the center line and is further axially away from the center line of the turbine housing. The effective center 164 of the two vortex chambers of the turbine housing is now at the radius R _D biased to the left of the center line of the dividing wall.

In order to produce an optimal asymmetric turbine housing, the inventors started in the same sized vortex chambers from the option of increasing one vortex chamber A / R or decreasing one vortex chamber from the same size vortex chambers. Several options regarding the size of the vortex chamber have been considered, up to the option of increasing one vortex chamber A / R and decreasing the outer vortex chamber A / R. The inventors have confirmed through experiments that the later solution, which is a 60/40 division of "A" -sections with 60% center side and 40% outside, preferably divides the mass flow with the restriction valve fully open. .

In all divided turbine housings, there is a cross flow ?? between the ends of the minimum diameter dividing wall and the ends of the turbine wheel. In order to minimize turbine wheel operation by actuation of rotating turbine wheel blades through the stationary tongue 26 of FIG. 4 at approximately the beginning of the dividing wall, the depth of the dividing wall is typically defined by the diameter of the turbine wheel stage. It does not extend adjacent to the turbine wheel blade stages more than the ratio of 120% to 150% of D ₃ ). This ratio of D _bc / D ₃ is usually determined by company design rules and technical goals. The diameter D _bc is known as the base circle. Since the impurities in the molten cast iron enter the stage of the turbine housing dividing wall, the ends of the turbine housing dividing wall are prone to cracking, so the innermost or minimum diameter of the dividing wall is typically at least 120% of the diameter of the turbine wheel stage. 150%. This ?? between the dividing wall and the turbine wheel enables cross-talk between the pulses of the exhaust flow as well as the cross flow of the exhaust gas between the two vortex chambers, the crosstalk initially being in the first position. This is the reason for having a split wall.

If the diameter of the turbine stage with the base circle is 120% of the diameter of the turbine wheel, the area is between 70% and 105% of the area of the two vortex chambers in a symmetrical configuration in section "A" of the turbine housing assembly of 5 A / Rs. There is a cross flow ?? In the case where the diameter of the turbine stage with the base circle is 150% of the diameter of the turbine wheel, the cross flow ?? is from 199% of the area of the two vortex chambers in a symmetrical configuration at the cross section "A" of the turbine housing assembly of 5 A / Rs. It has an area of 299%. This analysis shows that the membrane area can provide a very large cross sectional area for cross flow from one vortex chamber to another.

Since the membrane area is a function of both the turbine wheel diameter D ₃ and the minimum position D _bc of the dividing wall, the membrane area is variable due to the different values of D ₃ . 17 shows the membrane areas for an example of turbine wheel diameters produced from 64 to 94 mm. The film areas 133 are surrounded between the upper boundary line 131 and the lower boundary line 132 or are limited by the upper boundary line 131 and the lower boundary line 132. As expected, the range of film areas increases with increasing D ₃ . In FIG. 17, turbine wheel diameters D ₃ are shown at 123, and line 124 shows the trend of turbine wheel diameters D ₃ of the analyzed turbos. This chart also displays D _bc / D _{3 in the} range 1.25 to 1.35. Contains rain.

The inventors experimented with a 64mm turbine wheel, in a combination of asymmetrically configured 60/40 vortex chambers and limiting valves, as well as the ports of the dividing wall, as well as the foundation (determined by the difference excluding the area of D ₃ from the area of D _bc ). It has been found that the optimal crossflow area, including the area of the circle "membrane", is an area having an area of 289.6% of one, symmetrical, vortex compartment cross-sectional area of section "A" (ie half area of section "A"). . This is comparable to the typical cross flow area of a turbine housing of the same size without slots or ports and with the same D _bc / D ₃ ratio having a cross flow area of only 182.6% of the half area of section “A”. do.

As in the relationship between the membrane areas and D ₃ and D _bc , the total cross flow area 122 is affected by changes in the area of D ₃ and D _bc as well as one vortex chamber of section “A”. Cross flow areas 122 are surrounded by an upper boundary line 126 and a lower boundary line 127. The chart of FIG. 18 shows the relationship of the cross flow area and D ₃ 123 of the other turbine stages analyzed by the inventors.

To select the cross flow area, determine the value of diameter D ₃ of the turbine wheel in inches. One example is a 76 mm (2.992 ") turbine wheel diameter, shown as horizontal line 128. From the turbine wheel diameter, a vertical line 129 intersecting the turbine wheel diameter 123 connects the lower boundary line 127 and the upper boundary line 128. The cross flow area is shown by the vertical portion 130 of the vertical line 129 between the lower and upper boundary lines 127, 126.

The equation for generating the data points indicated in the charts shown in FIGS. 17 and 18 may be as follows:

As shown in FIG. 9, in low flow conditions, the pivoting valve member 72 is the center or bearing housing side that flows from the manifold to the turbine wheel 70 through a small vortex chamber 49 on the outside or outlet side. To restrict flow to the large vortex chamber 48.

The asymmetric turbine housing has a large left or center side vortex chamber 48 and a small right or outer vortex chamber 49 positioned axially with respect to the dividing wall 25. The flow restrictor, in this case the pivot valve member 72, is constrained to the mating surfaces of the manifold center plane foot 37 and the turbine housing foot 51. The inventors choose this configuration for cost and technical reasons, but the limiting device can be located in the central side exhaust manifold passage 34.

As shown in FIG. 10B, in high flow conditions, the rotary valve member 72 is not biased either in the large center side vortex chamber 48 or the small outer vortex chamber 49 to allow maximum flow to the turbine wheel. It is in the center position. In this high flow condition, the rotary valve member 72 of the flow restrictor is aligned with the dividing wall 25 of the turbine housing downstream of the turbine foot 51. In the minimum flow position (shown in dashed lines in FIG. 10B), the rotary valve member 72 restricts the exhaust flow to the large vortex chamber 48 so that the exhaust flow flows through the small vortex chamber 49, preferably. Is rotated toward the closed position about the axes 74, 78 of the device by the force applied to the actuating arm 73. The flow restrictor can be adjusted to any position between fully open or fully closed positions.

A cross-sectional view of this type of flow restrictor is shown in FIG. 10A. In this respect, in a preferred embodiment of the present invention, it can be seen that the flow restricting blade is made with two cylindrical bearing surfaces for rotation and an actuating arm 73 for position control. One side of the hole formed at the connection portion of the turbine housing foot 51 and the exhaust manifold foot 37 and accommodating the bearing surface is a blocked hole 77, while the other side 75 is an open hole. On the open orifice side, the piston ring 76 provides a gas seal as well as axial alignment for the flow restrictor. The placement of the actuating arm 73 can be optimized to meet design constraints.

The inventors find that, at both sea level and elevation, not only the back pressure as a function of engine speed and load, but also the ratio of boost-to-back pressure increases, making the flow restrictor of the exhaust system an ideal parameter. It became. When the rotatable flow restrictor is rotated toward the closed position, the turbine housing operates as if it were a smaller A / R turbine housing than present with the flow restrictor open. This increases the exhaust back pressure necessary for the EGR flow from the exhaust side of the engine to the intake side of the engine. Thus, the rotation of the flow restrictor can be used to make a pressure difference (from the exhaust side of the engine to the intake side of the engine) to assist the EGR flow from the exhaust side of the engine to the intake side of the engine.

In the first embodiment of the present invention, the effective mass flow to the turbine wheel is caused by the flow restrictor rotating about the point of the turbine housing inlet or foot, whereby the rotary valve member 72 of the flow restrictor in the open position is exhausted. It is controlled to be in line with the dividing wall 25 of the turbine housing, which minimizes the restriction of flow. As it is necessary to further limit the mass flow to the turbine wheel or reduce the mass flow, the pivot arm 73 is operated to rotate about its axes 74 and 78, such that the swirl valve member 72 has a large vortex chamber. This impedes the flow of exhaust gas to 48, which tunably reduces the mass flow to the turbine wheel.

In a variant of the first embodiment of the present invention, as shown in FIGS. 11A and 11B, the flow restrictor takes the form of a butterfly valve 80, which reduces the moment at the pivot arm 81. This allows for lower cost actuators with less force. In Fig. 11A, which is a sectional view of a first modification of the first embodiment of the present invention, the configuration of the bearing surfaces and the piston ring is the same as that of the first embodiment. In FIG. 11B, the rotation of the butterfly 77 about its axes 74, 84 at the center of the flow path to the central side vortex chamber 48 with a large rotational position in the case of a butterfly configuration results in an outer vortex An adjustable flow restriction is provided to the central side vortex chamber 48 which deflects flow into the chamber 49. For this variant of the first embodiment of the present invention, in the minimum flow restriction position, as shown in Fig. 11B, the butterfly valve is aligned side by side with the flow through the vortex chamber so that the stages of the butterfly valve are vortexed in the center side The seal 48 is closed or covered. The solution using a butterfly valve has the advantage of a low operating load since the moments on both sides of the rotation cancel each other out.

When the flow restrictor is in the partially open position, flow from the outer (small) vortex chamber 49 to the central (large) vortex chamber 48 reduces the length of the partition wall 25 or slots in the partition wall. By forming them.

 In general, in the commercial diesel market, for products where high speed travel can be expected, the split walls of the turbine housing are prone to cracking. The inventors have recognized the opportunity to mechanically minimize this tendency for cracks in the dividing wall by introducing pre-cast stress-relieving features into the dividing wall. FIG. 12B shows the turbine housing seen along the A-A cross section of FIG. 12A. This cross-section is generally performed to assess the condition of the turbine housing after a thermal cycling qualification test that causes the turbocharger to undergo extreme temperature cycles to measure the resistance of the turbocharger to cracks. . In FIG. 12B, the cracks 87 shown are typical of a commercial diesel type turbine housing in a split wall area.

The inventors believe that if "stress removal means" in the form of slots or ports are cast on the dividing wall, these ports not only minimize the propensity for cracking, but also the adjustable center from the unregulated outer vortex chamber in part to full closure of the restriction valve. It was assumed that it would provide a flow path to the side vortex chamber. Additional flow paths provide flow to the turbine wheel over a larger peripheral gap or area than is possible without slots or ports.

In the second embodiment of the present invention, as shown in FIG. 13, the effective mass flow to the turbine wheel is controlled by the flow restricting device of the asymmetric turbine housing with cross flow ports 88 formed in the dividing wall. In a second preferred embodiment of the invention, the area of the ports is surrounded by the leading radius 89, the trailing radius 90, the inner end circular portion 92 and the outer end spiral 91 of each port. All. The sum of the areas of the cross flow ports of the turbine housing is approximately equal to the area of the adjustable vortex chamber of section "A". What is important is not the shape of the ports, but the sum of the areas of the ports.

In a variant of the second embodiment of the invention, as shown in FIG. 14, the effective mass flow to the turbine wheel is controlled by the flow restriction device of the asymmetric turbine housing with cross flow slots 95 formed in the dividing wall. . In a second preferred embodiment of the invention, the area of the slots is at an inner boundary by the leading radius 98, the trailing radius 99, the outer spiral 91 and the base circle 71 of each port. Surrounded by The sum of the areas of the cross flow slots of the turbine housing is approximately equal to the area of the adjustable vortex chamber of section "A". What is important is not the shape of the areas of the slots but the sum of the areas of the slots and the cross flow area inside the dividing wall stage. In a second preferred embodiment of the present invention, the outer boundary 97 of the slot is characterized by a keyhole at the outer end of the slot in order to increase the stress and minimize the propensity of the slot in which the initial crack occurs. Can lose.

As in the basic twin flow turbine housing, multiple flow turbine housings with vortex compartment partition walls parallel to the axial surfaces, ie turbocharger axes, rather than radial surfaces, are not common. The inventors have the opportunity to use similar logic for multiple flow turbine housings with asymmetric vortex compartment areas involving flow restrictors to more cost effectively increase the flow range of the turbine stage with this type of turbine housing. Recognized.

In a third embodiment of the invention, a triple flow turbine housing such as shown in FIG. 15A is preferably used. The two axial vortex compartment dividing walls 103, 104 are assembled in the turbine housing such that the ratio of flows from outside to the interior passing through the unrestricted adjacent vortex compartments is approximately 70% to 20% to 10%. These ratios can be changed as needed. The ratio of the flows is only important in that the adjustable vortex chamber is such that the sum of the open areas is equal to the area of the adjustable restriction valve. Flow restriction valves are provided. The flow restriction valve is rotated about the turbine housing inlet or foot point so that in the open position the blade 89 of the flow restriction device is flush with the turbine housing outer vortex wall wall to minimize the restriction of exhaust flow. Since it is necessary to further restrict or reduce the mass flow to the turbine wheel, the pivot arm 73 is operated to rotate about an axis so that the blade 89 first of the exhaust gas to the outer vortex chamber (s), then to the central vortex chamber. Try to interrupt the flow. In this way, the effective mass flow to the turbine wheel is controlled through a flow restrictor that can controllably reduce the mass flow to the turbine wheel.

In a variant of the third embodiment of the invention, the dividing walls 106, 107 have a slot 108 such that flow from the outer vortex chambers reaches the inner vortex chambers and reaches the turbine wheel 70. do. Slots 108 also allow for mass flow regulation and allow for a more constant and desirable flow distribution to the turbine wheel.

Claims (14)

As a variable flow turbocharger,
(a) A turbine inlet is provided in the turbine housing foot 51 that engages the exhaust manifold to receive the exhaust gas flow, and a turbine outlet 52 is provided between the turbine housing foot and the outlet. A turbine housing 2 provided;
(b) a turbine impeller (70) having a plurality of blades located within the turbine housing for receiving exhaust gas flow from the vortex chamber;
(c) at least one vortex compartment dividing wall for dividing the vortex compartment into a large vortex chamber 48 and a small vortex chamber 49; And
(d) a variable flow turbocharger comprising exhaust flow control valve means adapted to control the degree of blockage of exhaust gas flowing into the large vortex chamber. The method of claim 1,
And the exhaust flow control valve means is variably adjustable. The method of claim 1,
And the vortex compartment dividing wall comprises apertures, and exhaust flow can pass between the large vortex compartment and the small vortex compartment through the apertures. The method of claim 1,
And the dividing wall is an axial dividing wall. The method of claim 1,
And the dividing wall is a radial dividing wall. The method of claim 1,
Wherein the volume of the large vortex compartment comprises at least about 55% of the divided vortex compartment space and the volume of the small vortex compartment comprises up to about 45% of the divided vortex compartment space. The method of claim 1,
Wherein the volume of the large vortex chamber comprises about 55-65% of the divided vortex chamber and the volume of the small vortex chamber comprises about 45-35% of the divided vortex chamber. The method of claim 1,
A variable flow turbocharger comprising at least two axial or radial dividing walls. The method of claim 1,
The dividing wall is radial, the small vortex chamber (49) is on the outlet side and the large vortex chamber (48) is on the opposite side of the outlet side. The method of claim 1,
The exhaust flow control valve means includes a rotary valve member 72 configured to rotate between an open position in which flow to the large vortex chamber 48 is not limited and a closed position in which flow to the large vortex chamber 48 is effectively blocked. Variable flow turbocharger, characterized in that. An internal combustion engine comprising an exhaust manifold and having a variable flow turbocharger fluidly coupled with the exhaust manifold, the variable displacement turbocharger is:
(a) a turbine housing 2 having a turbine inlet, a turbine outlet 52 and a vortex chamber;
(b) a turbine impeller positioned in the turbine housing and having a plurality of blades;
(c) at least one vortex compartment dividing wall extending near the turbine inlet to divide the vortex compartment into a large vortex chamber 48 and a small vortex chamber 49;
(d) a divided exhaust manifold (36) comprising a first exhaust inlet passage for introducing exhaust gas into the large vortex chamber and a second exhaust inlet passage for introducing exhaust gas into the small vortex chamber; And
(e) internal combustion engines comprising exhaust flow control valve means located in at least the first exhaust inlet passage to control the degree of blockage of the exhaust gas flowing into the first exhaust inlet passage. 12. The method of claim 11,
Said exhaust flow control valve means being located in a turbine housing. 12. The method of claim 11,
Said exhaust flow control valve means being located in an exhaust manifold associated with said turbine housing. 12. The method of claim 11,
And said exhaust flow control valve means is located between the exhaust manifold and the turbine housing.

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