DE102015103615A1 - centrifugal compressors - Google Patents

Abstract

A centrifugal compressor having a housing and a supersonic compressor impeller rotatably supported within the housing, wherein the housing forms an inlet and an outlet arranged such that the supersonic compressor impeller is flowed in the axial direction by a gas to be compressed and in the radial direction is formed, and in which between the supersonic compressor impeller and the outlet an annular diffuser space (40) delimited by opposite housing walls is formed, is characterized in that the housing walls in a first, emanating from the supersonic compressor impeller section (48 ) of the diffuser space (40) and in a second section adjoining the first section (48) extend parallel to one another or move away from one another.

Description

The invention relates to a centrifugal compressor, in particular a centrifugal compressor for charging an internal combustion engine. The internal combustion engine can be provided in particular for driving a motor vehicle.

The range of passenger cars offered by car manufacturers has been showing a trend towards engine performance in recent decades. On the one hand, this is due to the increased driving power requirements of customers. On the other hand, the increased vehicle weight due to higher safety and comfort requirements require performance-enhanced engines. The increase in power is usually not on the classic way of a maximum speed and / or displacement increase, but by a motor charging.

Especially in the last ten years, the engine charging has prevailed as part of the so-called "downsizing", but also as a measure to reduce fuel consumption. Downsizing reduces the displacement of the internal combustion engine with the goal of reducing absolute fuel consumption. However, the displacement reduction tends to result in a power decrease, which can be compensated by the use of a charge or even overcompensated.

An engine supercharging can be implemented by means of a so-called mechanical supercharging or by means of an exhaust gas turbocharging. The main distinguishing feature of both variants is the type of drive of the compressor. During turbocharging, the compressor is driven by a turbine integrated into the exhaust system of the engine, which in turn is driven by the exhaust gas mass flow. During mechanical charging, the compressor is driven by the engine directly, usually by the crankshaft by means of a belt drive. The compressor types used for the two different supercharging variants may also differ. While only turbochargers are used for turbocharging, superchargers are also used for mechanical turbocharging. As a flow loader, only centrifugal compressors are used in the passenger car sector; Axial compressor, however, for reasons of poorer efficiency not.

A radial compressor is in contrast to a turbine to the group of machines. It transfers energy to the gas flowing through it, which flows axially into the compressor and flows radially outward. For this purpose, a radial compressor comprises a housing which forms an inlet and an outlet. Within a flow space of the housing, a compressor impeller is rotatably mounted. This subdivides the flow space into the vacuum chamber comprising the inlet as well as into the overpressure space comprising the outlet. The overpressure space is composed of the diffuser adjoining the compressor impeller in the flow direction of the gas flow, the adjoining so-called spiral and the stub diffuser forming the outlet. The purpose of the spiral is to receive the flow at the circumference of the diffuser and then to pass it on to the nozzle diffuser, which further increases the static pressure.

The purpose of the diffuser is to use the kinetic energy of the gas flow generated by the compressor impeller for the static pressure build-up. For this purpose, both a bladed and a bladed diffuser can be used. Bladed diffusers have the advantage of a higher efficiency compared to bladed diffusers, for example at an outflow angle of less than 30 °. This is justified by the longer flow path, which travels a gas in a bladeless diffuser. Because of the swirl and mass conservation, the fluid in a parallel-walled diffuser on a logarithmic spiral flows outward only gradually. In contrast, the gas flow is given in a bladed diffuser the way, whereby the friction losses decrease due to the shorter distance. However, the blading of a diffuser has the disadvantage that the width of the compressor map is reduced. The right side of the map, which is called Stopfgrenze at the maximum enforceable mass flow, shifts to the left, since the flow cross-section in the diffuser is subject to a reduction by the blading. The surge limit also shifts. It starts at larger mass flows. Namely, the blading produces a higher flow resistance, whereby the gas flow is retarded stronger and thus earlier dissolves.

In aircraft engines, axial compressors are used. These differ from the centrifugal compressors in addition to the design by a special mode of operation. So they work in contrast to the centrifugal compressors in trans- and supersonic. Axial compressors with a supersonic mode of operation have been used in aircraft engines since 1968. This technology increased the overall pressure ratios per compressor stage, which in addition to the increase in power densities, a reduction in the number of compressor stages was possible.

Supersonic axial compressors generate a relative supersonic flow at the impeller inlet at sufficient speed by transition from absolute to relative reference system. Likewise, a supersonic flow can occur at the impeller outlet by changing from the relative to the absolute system. Compressors with this mode of operation are referred to as Transsonikverdichter because they only reach supersonic in the outer section. In the hub section, however, the flow remains subsonic due to the lower peripheral speed of the compressor impeller.

Also in the centrifugal compressor can be generated equivalent to the axial Transsonikverdichtern by increasing the speed of a supersonic flow at the impeller inlet and / or outlet. However, this type of supersonic production requires very high speeds of the compressor impeller, which is higher than the load limit conventional compressor used in the field of automotive technology. Thus, the supersonic mode of operation of the axial compressor used in aircraft engines is basically transferable to centrifugal compressor. However, radial compressors with this technology can not be used with the currently used materials (mainly aluminum alloys).

The US 3,010,642 discloses an alternative possibility for realizing a supersonic function in radial compressors, which is thus not based on an increase in the speed. The radial compressor described therein is based on a special design of the compressor impeller. This differs from the compressor impellers otherwise used in centrifugal compressors in the shape of the impeller passages formed between the impeller vanes. This special form of impeller ducts enables supersonic generation in the relative system and thus acceleration of the meridional component of the absolute velocity of the gas flow to supersonic. The designated compressor impeller has comparable to a Laval nozzle designed impeller channels, whereby the meridian regardless of the speed of the compressor impeller can reach supersonic. The speed rather only affects the pressure curve in the impeller channels. In contrast to a Laval nozzle, however, the design of the impeller channels must also take into account the centrifugal forces generated by the rotation of the impeller. In addition, occurs in these impeller channels compared to a gas flow in a Laval nozzle due to the centrifugal forces less pressure reduction.

That from the US 3,010,642 known compressor impeller has a relatively low impeller pressure ratio of 0.91 compared to conventional impellers. This is due to the fact that conventional wheels already delay the flow. The delay is thus not only in the diffuser of a corresponding centrifugal compressor. That from the US 3,010,642 however, known compressor impeller accelerates the gas flow. The delay of the gas flow thus takes place only in the diffuser of the centrifugal compressor.

From the DE 1053714 is a radial compressor with a bladed diffuser known, wherein a section comprising the vanes of the diffuser is preceded by an unspotted guide channel. This guide channel is intended to delay the gas flow from a supersonic speed to a subsonic speed. This is intended to prevent the occurrence of compression shock waves in the transition from supersonic to subsonic speed and consequently a high degree of efficiency is made possible.

Furthermore, from the US 3,604,818 a radial compressor for a gas turbine, in which a compressor impeller surrounding the diffuser of a plurality of each tangent to the outer periphery of the compressor impeller aligned diffuser channels composed. The diffuser channels have a rectangular cross-section, which increases, based on the flow direction of the compressed gas. It is stated that the gas flow is intended to discharge the compressor impeller at supersonic speed.

From the JP 3260399 A a radial compressor with an annular diffuser is known, in which the two walls delimiting the diffuser linearly approximate in a first section adjacent to the compressor impeller so that the flow cross section remains constant while they are parallel in a second section, due to Ring shape of the diffuser leads to an increase in the flow cross-section with increasing radial distance to the compressor impeller.

And finally is out of the DE-OS 1 628 227 a radial compressor with a blade-less diffuser, wherein in a portion of the diffuser a cross-sectional reduction of the free flow space is provided by an arcuate recess of one or both of the housing walls forming the diffuser.

Based on this prior art, the invention has the object to improve the operation of a centrifugal compressor and in particular the achievable by this pressure increase of to increase promotional gas, without the need for a larger dimensioning of the radial compressor or an increase in the drive power for the centrifugal compressor.

This object is achieved by means of a centrifugal compressor according to the patent claim 1. Advantageous embodiments thereof are subject matter of the dependent claims and will become apparent from the following description of the invention.

A basic idea of the invention lies in the fact that the stated object can be achieved if a gas flow flowing away from a compressor impeller at supersonic speed is delayed as little as possible to subsonic sound.

Another basic idea of the invention is that such a low-loss delay of the gas flow can be achieved if a working according to the reverse Laval nozzle principle shock diffuser, as it is known from axial compressors, modified and adapted to the particular conditions for a diffuser of a centrifugal compressor becomes. It was recognized that such a diffuser to delay the gas flow not only in the supersonic range, so it should not only act as a supersonic diffuser, but also the gas flow must be slowed down to subsonic. It was further recognized that this compared to that diffuser, which in the radial compressor according to the US 3,010,642 is used, at least the static efficiency and / or the total pressure loss can be improved.

These principles and findings of the invention can be generally implemented in that in a centrifugal compressor comprising at least one housing and a supersonic compressor impeller rotatably mounted within the housing, the housing forming an inlet and an outlet arranged such that the supersonic compressor impeller would be impinged by a gas to be compressed in the axial direction and flowed away in the radial direction, and between the supersonic compressor impeller and the outlet would be a (partially) annular, at least partially encircling housing walls surrounding the supersonic compressor impeller limited diffuser space is formed, the housing walls in a first, emanating from the supersonic compressor impeller portion of the diffuser space and in a second, to the first portion (in the radial direction with respect to a rotational axis of the supersonic compressor impeller ) subsequent section parallel to each other or from each other. In this case, the parallel or mutually removing radial course of the housing walls in the second section is always connected to an enlargement of the flow cross-section formed by the latter due to the annular configuration of the diffuser space.

A "supersonic compressor impeller" is understood according to the invention to mean a compressor impeller that is structurally designed in such a way that the gas flow and, in particular, the meridional component of the absolute velocity of the gas flow at the time of outflow of the compressor impeller at supersonic speed are possible.

For this purpose, it can preferably be provided that the flow cross-section of at least one impeller channel formed between two impeller blades and preferably the flow cross-sections of several or all of these impeller channels in at least one subsequent to a Laufradeintritt section and preferably over the entire length of the / the impeller channel (s) in the flow direction ( preferably continuously), as is basically the case of US 3,010,642 is known. However, a conventionally designed compressor impeller, which is basically characterized in that the flow cross sections of the impeller channels continuously increase from the impeller inlet in at least one section and in particular over the entire impeller channel length, can be designed as supersonic compressor impeller, if such a constructive so is interpreted that depending on the geometric dimensioning so high speeds are possible that the desired supersonic velocity of the gas flow is achieved. It seems possible that the meridional component of the gas flow may also reach supersonic speed, with permanent acceleration of the meridional component to supersonic only appear to be feasible if the total cross-sectional area at a location of the impeller channel (s) is in the range of Ma = 1. In such a constructive design of a conventional compressor impeller, in particular the choice of material can play a relevant role, because by the use of high-strength (especially metallic) materials which simultaneously have a relatively low density, higher maximum speeds for compressor impellers can be achieved than in the classic Construction materials (especially aluminum alloys) for compressor wheels is the case.

Preferably, it can be provided that the approach of the housing walls in the first section and / or a distance of the housing walls from each other in the second section is continuous and in particular linear. This may have an advantageous effect on the gas flow in this / this section (s). In addition, this can bring advantages in terms of manufacturability of the centrifugal compressor.

In a further preferred embodiment of the radial compressor according to the invention, it can be provided that the flow cross-section of the diffuser space decreases over the radial height of the first section. For this purpose, the approach of the housing walls in this first section should therefore be so large that it overcompensates the increase in the flow cross-section, which results from the (partial) ring shape of the diffuser space in conjunction with an increasing distance from the rotational axis of the supersonic compressor impeller , It can thereby be achieved that the first section of the diffuser space acts as a supersonic diffuser and can prevent the formation of compaction shocks.

The approach of the housing walls in the first section and / or a distance of the housing walls from each other in the second section may preferably be formed on both sides, so that the opposite housing walls each have a course in the radial direction, which is closer to a radial plane lying between them leads. This can have a particularly advantageous effect on the realization of the highest possible efficiency and / or the lowest possible total pressure loss in the diffuser.

In a further preferred embodiment of the radial compressor according to the invention can be provided that the transition between the first portion and the second portion is arranged at least half the radial height of the diffuser space. This can often be ensured that the gas flow is delayed until reaching the transition to subsonic, which can be beneficial to the efficiency of the compressor and / or the total pressure loss.

Preferably, it can be provided that the transition between the first portion and the second portion is angled. This can have an advantageous effect on the desired flow guidance within the diffuser space. Furthermore, this can simplify the manufacturability of the radial compressor according to the invention. The design with a (rounded with a constant or varying radius) transition is also possible.

In a further preferred embodiment of the radial compressor according to the invention can be provided that the diffuser space is unencumbered, because this can lead to the realization of particularly high efficiencies and / or very low total pressure losses in combination with the inventive design of the diffuser.

The invention further relates to an exhaust gas turbocharger with a turbine and a radial compressor according to the invention. In this case, the turbine forms a housing and a turbine runner, wherein the turbine runner is connected directly, for example via a shaft, or indirectly via a gear to the compressor wheel in order to drive this can.

Furthermore, the invention relates to an internal combustion engine with an internal combustion engine, a fresh gas train, an exhaust system and a radial compressor according to the invention or an exhaust gas turbocharger according to the invention. Fresh gas, which has been previously compressed by means of the radial compressor according to the invention, can be supplied to the internal combustion engine via the fresh gas train, which gas can then be burned together with fuel in combustion chambers of the internal combustion engine. The exhaust gas generated in this case can be discharged via the exhaust line of the internal combustion engine, which optionally flows through the turbine integrated in the exhaust line of an exhaust gas turbocharger according to the invention, while the turbine wheel rotates, which in turn the compressor impeller of the compressor is driven in rotation to cause the compression of the fresh gas ,

Furthermore, the invention relates to a motor vehicle, in particular a wheel-based motor vehicle (preferably a car or truck), with an internal combustion engine according to the invention. In this case, the internal combustion engine can be provided in particular for (direct or indirect) provision of the drive power for the motor vehicle.

The indefinite articles ("a", "an", "an" and "an"), in particular in the patent claims and in the specification generally explaining the claims, are as such and not as numerical words to understand. Corresponding to this concretized components are thus to be understood that they are present at least once and may be present more than once.

The present invention will be explained in more detail with reference to an embodiment shown in the drawings. In the drawings shows:

1 in a schematic representation of an internal combustion engine with an exhaust gas turbocharger comprising a radial compressor according to the invention;

2 in a schematic representation of an internal combustion engine with a mechanically driven radial compressor according to the invention;

3 a radial compressor according to the invention in a longitudinal section;

4 : the centrifugal compressor according to the 3 in a partially sectioned view from the front;

5 a first possible embodiment of the diffuser chamber of the radial compressor according to the 3 and 4 in an enlarged view of the section marked V in the 3 ;

6 a second possible embodiment of the diffuser space of the radial compressor according to the 3 and 4 in an enlarged view of the marked with VI section of 3 ;

7 a schematic comparison of diffusers and nozzles for subsonic and supersonic flows;

8th schematically the formation of compression shocks in a shock diffuser;

9 FIG. 2 schematically shows the various components of the flow velocity of a gas flow in the area of the entry and exit of a compressor impeller; FIG.

10 : the course of the Mach number, the density and the total pressure loss in a first variant of a diffuser according to the 5 ;

11 the course of the Mach number, the density and the total pressure loss in a second variant of a diffuser according to the 5 ;

12 : the course of the Mach number, the density and the total pressure loss in a third variant of a diffuser according to the 5 ;

13 : the course of the Mach number, the density and the total pressure loss in a first variant of a diffuser according to the 6 ;

14 the course of the Mach number, the density and the total pressure loss in a second variant of a diffuser according to the 6 ; and

15 : the course of the Mach number, the density and the total pressure loss in a third variant of a diffuser according to the 6 ,

The 1 shows a schematic representation of an internal combustion engine with an internal combustion engine 10 that has a plurality of cylinders 12 formed. The cylinders 12 along with piston reciprocated therein and a cylinder head (not visible) define combustion chambers in which fresh gas (air) is burnt together with fuel, causing the pistons to be cycled up and down. This movement of the piston is in a known manner to a in the 1 not shown crankshaft 14 (see. 2 ) and thus driven rotating. The fresh gas becomes the internal combustion engine 10 supplied via a fresh gas train, in which a radial compressor according to the invention 16 is integrated as part of an exhaust gas turbocharger. The fresh gas is by means of the centrifugal compressor 16 compressed, then in a charge air cooler 18 cooled and controlled by a throttle 20 fed to the combustion chambers.

The drive of the centrifugal compressor 16 takes place by means of a turbine 22 , which is integrated in an exhaust line of the internal combustion engine. Exhaust gas, the combustion of the fuel-fresh gas mixture in the Combustion chambers of the internal combustion engine 10 is created via the exhaust system of the internal combustion engine 10 dissipates and flows through the turbine 22 , This leads in a known manner to the rotating drive of a turbine runner, which in turn via a shaft 24 rotatably with a compressor impeller of the centrifugal compressor 16 connected is. The rotating drive of the turbine runner is thus transferred to the compressor runner. To operate the internal combustion engine 10 Limiting the pressure build-up in the fresh gas line with high speeds and loads is the turbine 22 in a known manner by means of a so-called wastegate 26 bypassed.

In the internal combustion engine according to the 2 becomes the integrated in the fresh gas train radial compressor 16 mechanically, ie directly from the crankshaft 14 of the internal combustion engine 10 by means of a belt drive 28 driven. Since in this case the speed of the compressor impeller is proportional to the speed of the crankshaft 14 of the internal combustion engine 10 is, there is the possibility of pressure build-up by means of the centrifugal compressor 16 at high speeds of the crankshaft 14 by means of a compressor bypass 30 limit in a known manner.

The 3 and 4 show a possible embodiment of a radial compressor according to the invention 16 , as used, for example, in internal combustion engines according to the 1 and 2 can be used. This radial compressor 16 includes a housing 32 and a rotatable within the housing 32 stored compressor impeller 34 , The housing 32 forms an inlet 36 and an outlet (not visible). The outlet is formed by a nozzle diffuser (not visible), which is tangent to the so-called spiral 38 designated, from the housing 32 limited annulus going off the compressor impeller 34 radially surrounds. The transition from the compressor impeller 34 to the spiral 38 one of the compressor impeller 34 radially surrounding, from the housing 32 limited diffuser space 40 formed compared to the spiral 38 a significantly smaller flow cross-section for the flow of the fresh gas to be compressed and also has a much smaller volume. This diffuser room 40 serves to make the compressor impeller 34 To decelerate at supersonic velocity flowing gas flow to subsonic and simultaneously increase the pressure of the fresh gas to the desired compression effect of the centrifugal compressor 16 to reach.

At the compressor impeller 34 of the centrifugal compressor according to the 3 and 4 it is a supersonic compressor impeller 34 because this is due to its design capable of the radial compressor 16 passing gas flow, and particularly the meridian component c _m of the absolute velocity of this gas flow independently of the speed n of the compressor impeller 34 to accelerate to supersonic. For this purpose, it is provided that the flow cross sections of impeller passages (located perpendicular to the main flow direction of the fresh gas flow) 42 coming from neighboring impeller blades 44 of the compressor impeller 34 as well as from the case 32 be limited, starting from an entry of the compressor impeller 34 towards an exit of the compressor impeller 34 reduce, as a result of increasing with radial distance from the axis of rotation 46 of the compressor impeller 34 increasing distance between adjacent impeller blades 44 one over the length of the impeller channels 42 (in the flow direction) so greatly reducing depth of the impeller channels 42 is provided that tends to increase with increasing radial distance from the axis of rotation 46 successive enlargement of the flow cross sections of the impeller channels 42 is overcompensated.

Such a configuration of a compressor impeller 34 and the fact that with such a compressor impeller 34 an acceleration of the meridional component c _m one by such a radial compressor 16 guided gas flow can take place on supersonic, is basically from the US 3,010,642 known. This desired supersonic effect of the compressor impeller 34 is based on the principle of a Laval nozzle, which is the only flow-conducting component, by means of which a gas flow can be accelerated from supersonic to supersonic without the supply of potential energy. Such a Laval nozzle is basically characterized by a first section in which the flow cross-section is reduced. In a subsequent to the first portion in the flow direction of the gas flow second portion of the flow cross-section increases again. In order to be able to achieve an acceleration of the gas flow to supersonic, the transition between these two sections, in which the flow cross section is smallest, must be located where the gas flow exceeds the speed of sound.

For the compressor impeller 34 according to the 3 and 4 it is not necessary that the impeller channels 42 after a first, tapered section re-expand to a flow of the compressor impeller 34 with supersonic speed also with regard to the meridional component c _{m to} achieve. This is because, unlike a Laval nozzle, the potential energy of the gas flow does not remain substantially the same, but considerably by the rotating drive of the compressor impeller 34 and the acceleration of the gas flow caused thereby is increased.

If a supersonic flow is compared with a subsonic flow, it must always be stated that these behave in opposite ways to one another. For example, a subsonic flow in a straight tube of constant cross-section experiences a decrease in static pressure due to friction due to friction, thereby causing an increase in speed and a drop in temperature. If, on the other hand, this pipe flows through supersonic, the flow velocity drops and the static pressure and the temperature increase. Furthermore, a supersonic flow for their deceleration and acceleration requires a different diffuser and nozzle shape than a subsonic flow. A subsonic diffuser, ie a diffuser for subsonic flows, has a flow cross-sectional widening and a subsonic jet has a flow cross-sectional taper, such as 7 shows. Conversely, for a supersonic diffuser, a flow cross-sectional taper and for supersonic nozzle a flow cross-sectional widening is required (cf. 7 ). Thus, a supersonic diffuser for supersonic flow acts as a nozzle and vice versa.

The reason for these different geometric shapes of diffusers or nozzles for supersonic or subsonic sound is the behavior of the air density ρ, since this increases more strongly in a supersonic diffuser, for example, than the velocity v decreases. Consequently, according to the equation of continuity, the flow cross-section (the flow cross-sectional area) A must be reduced: A = ṁ / ρ · ν

If it is intended to delay an air flow from supersonic to subsonic to optimum efficiency, are in the 7 assembled so that the air flow first passes through the supersonic and then the subsonic diffuser (shock diffuser). Equivalently, for the acceleration of a flow from subsonic to supersonic, it is necessary to string together the subsonic and supersonic nozzles (Laval nozzle). However, a deceleration to subsonic and an acceleration to supersonic can only be achieved if the flow in the narrowest cross-section of the geometry, the critical flow cross-section A _critical , the critical pressure ratio π _critical = 1.89 and thus reaches Ma = 1:

In a Laval nozzle, the acceleration process is continuous; There is no compression shock. However, the outlet cross-section needs to be adapted to the backpressure. In contrast, in a shock diffuser at least one compression shock always occurs. The process is thus unsteady and therefore very lossy. To minimize the losses, it is therefore advisable to keep the vertical compression shock, which is necessary at least in the case of swirl-free inflow to achieve the subsonic speed, small. This is achieved by several oblique compression strokes, which reduce the Mach number as much as possible before the formation of the vertical impact. This is schematically in the 8th clarified.

The generation of a supersonic flow in a centrifugal compressor is basically possible by increasing the speed. If this is sufficiently large, supersonic speed is achieved in the impeller inlet and outlet. An increase in the speed n of the compressor wheel leads namely to an increase in the peripheral speed at the impeller inlet (u ₁ ) and outlet (u ₂ ). If the vector u ₁ now with the absolute inflow velocity c ₁ according to the velocity triangle in the 9 added, results in sufficient speed increase supersonic relative velocity w ₁ in the impeller inlet. If the vector sum of relative velocity w ₂ and peripheral velocity u _{2 is} sufficiently large, the absolute velocity c _{2 reaches} supersonic. The emergence of supersonic at the impeller inlet and outlet thus results from the change to another reference system. At the impeller inlet, the transition from an absolute to a relative reference system and at the impeller outlet from a relative to an absolute reference system is feasible. These supersonic gas flows have in common that, although the relative and / or absolute component of the gas flow is accelerated to supersonic by vector addition or by changing the reference system, the meridional component c _m reaches a maximum of Ma = 1. It is namely is the same for both the relative and absolute reference systems. For this reason, the geometry of the impeller channels must usually be adapted to the density change of the gas flow (at least if the flow of a compressor impeller is considered isolated). This means that the shape of the impeller channels should be designed according to a Laval nozzle. Thus, the meridional component c _{m of} the absolute velocity of the gas flow can be accelerated to supersonic speed, since only a Laval nozzle of a gas flow without reference system change allows exceeding the speed of sound.

The compressor impeller 34 of the centrifugal compressor 16 according to the 3 and 4 uses this and sees a corresponding design of the impeller channels 42 in the form of a Laval nozzle, whereby the meridional component c _m regardless of the rotational speed n of the compressor impeller 34 Can reach supersonic speed. The speed n only influences the pressure curve in the impeller channels 42 , Unlike a Laval nozzle must be in the design of the impeller channels 42 but also by the rotation of the compressor wheel 34 centrifugal forces are taken into account. In addition, occurs in such impeller ducts 42 compared to a Laval nozzle due to the centrifugal forces a less severe pressure reduction.

While the diffuser space in the centrifugal compressor according to the US 3,010,642 through a parallel or up to the spiral 38 is characterized approximate course of these limiting housing walls, is in the radial compressor according to the invention 16 provided that such an approximation only in a first, from the compressor impeller 34 radially outgoing section 48 is provided while in one is attached to the first section 48 subsequent second section 50 the the diffuser room 40 bordering housing walls run parallel (see FIG 6 ) or move away from each other again (see embodiment according to the 5 ). Here is the approximation of the housing walls in the first section 48 chosen so large in both embodiments, that is characterized by the increasing radial distance from the axis of rotation 46 resulting enlargement of the diffuser space 40 trained flow cross-section is overcompensated and thus reduces the flow cross-section (linear). In the second section 50 increases this flow cross section in both embodiments again (linear), which in the embodiment according to the 6 exclusively on the increasing radial distance to the axis of rotation 46 and in the embodiment according to the 5 in addition to the (with increasing radial distance from the axis of rotation) increasing distance of the diffuser space 40 limiting housing walls in this second section 50 is justified. It is envisaged that both the approach of the housing walls in the respective first section 48 both embodiments as well as the distance from each other in the second section 50 the embodiment according to the 5 is linear. The transition between the two sections is thereby angled.

Due to the inventive design of the diffuser space 40 a radial compressor 16 can with respect to the centrifugal compressor according to the US 3,010,642 a significant increase of the isentropic (diffuser) efficiency η _{D, s} and a reduction in the total pressure loss Δp _Tot can be achieved. These advantages of the inventive design of the diffuser space could be demonstrated by means of flow simulations. Among other things, six variants of diffuser spaces according to the invention were investigated; three each per embodiment according to the 5 and the 6 ,

• – Eintrittsdurchmesser des Diffusors: d₃ = 200 mm;
• – Austrittsdurchmesser des Diffusors: d₄ = 320 mm
• – Einlassbreite: x₀ = 4,2 mm.

The underlying the flow simulations diffuser spaces 40 had the following matching dimensions:

• - inlet diameter of the diffuser: d ₃ = 200 mm;
• - Discharge diameter of the diffuser: d ₄ = 320 mm
• - inlet width: x ₀ = 4.2 mm.

• – Auslassbreite: x₃ = 4,2 mm.

The three variants of the design of a diffuser room 40 according to the 5 had the following matching dimension:

• - outlet width: x ₃ = 4.2 mm.

• – Breite im Übergang zwischen dem ersten und dem zweiten Abschnitt: x_min = 3,2 mm
• – Durchmesser des Übergangs zwischen dem ersten und dem zweiten Abschnitt: d₅ = 220 mm.

The first variant of the embodiment of a diffuser room 40 according to the 5 also had the following dimensions:

• - Width in the transition between the first and the second section: x _min = 3.2 mm
• - Diameter of the transition between the first and the second section: d ₅ = 220 mm.

• – Breite im Übergang zwischen dem ersten und dem zweiten Abschnitt: x_min = 3,2 mm
• – Durchmesser des Übergangs zwischen dem ersten und dem zweiten Abschnitt: d₅ = 300 mm.

The second variant of the embodiment of a diffuser room 40 according to the 5 also had the following dimensions:

• - Width in the transition between the first and the second section: x _min = 3.2 mm
• - Diameter of the transition between the first and the second section: d ₅ = 300 mm.

• – Breite im Übergang zwischen dem ersten und dem zweiten Abschnitt: x_min = 2,2 mm
• – Durchmesser des Übergangs zwischen dem ersten und dem zweiten Abschnitt: d₅ = 300 mm.

The third variant of the embodiment of a diffuser room 40 according to the 5 also had the following dimensions:

• - Width in the transition between the first and the second section: x _min = 2.2 mm
• - Diameter of the transition between the first and the second section: d ₅ = 300 mm.

• – Auslassbreite und Breite im Übergang zwischen dem ersten und dem zweiten Abschnitt: x_min = 3,2 mm.
• – Durchmesser des Übergangs zwischen dem ersten und dem zweiten Abschnitt: d₅ = 220 mm.

The first variant of the embodiment of a diffuser room 40 according to the 6 also had the following dimensions:

• - outlet width and width in the transition between the first and the second section: x _min = 3.2 mm.
• - Diameter of the transition between the first and the second section: d ₅ = 220 mm.

• – Auslassbreite und Breite im Übergang zwischen dem ersten und dem zweiten Abschnitt: x_min = 3,2 mm.
• – Durchmesser des Übergangs zwischen dem ersten und dem zweiten Abschnitt: d₅ = 300 mm.

The second variant of the embodiment of a diffuser room 40 according to the 6 also had the following dimensions:

• - outlet width and width in the transition between the first and the second section: x _min = 3.2 mm.
• - Diameter of the transition between the first and the second section: d ₅ = 300 mm.

• – Auslassbreite und Breite im Übergang zwischen dem ersten und dem zweiten Abschnitt: x_min = 2,3 mm.
• – Durchmesser des Übergangs zwischen dem ersten und dem zweiten Abschnitt: d₅ = 300 mm.

The third variant of the embodiment of a diffuser room 40 according to the 6 also had the following dimensions:

• - outlet width and width in the transition between the first and the second section: x _min = 2.3 mm.
• - Diameter of the transition between the first and the second section: d ₅ = 300 mm.

The 10 and 11 show the simulation results to the first and second variants of the embodiment of a diffuser space according to the invention 40 according to the 5 ie with a width in the transition between the first and the second section (hereinafter referred to as "constriction") of 3.2 mm and a radial position of the constriction x _min at a diameter of 220 mm ( 10 ) or 300 mm ( 11 ). It can be seen that at a radius of about 125 mm (ie a diameter of about 250 mm) beginning portions of both flow constrictions are comparable. These areas are similar to an overexpanded supersonic jet. In contrast, the gas flows differ significantly up to this range. The diffuser room 40 with the constriction x _min at the diameter of 300 mm according to the 11 is subject to a change of acceleration zones and local delays due to compression shocks. At the diffuser room 40 with the constriction x _min at the diameter of 220 mm (cf. 10 ), the gas flow is initially delayed by the flow cross-sectional taper, but then accelerated again due to the increase in flow cross-section over the radius. The following free jet area is compared to the diffuser room 40 according to the 11 not parallel between diffuser floor and ceiling. Due to the separation behind the impeller transition, the gas flow initially experiences a certain deflection in the direction of the diffuser bottom. As a result, the gas flow settles on the diffuser ceiling due to the acceleration or is deflected in the direction of the diffuser cover, as a result of which it undergoes a deflection again at the detachment point. The in this variant of a diffuser according to the invention space 40 Flow separation occurring behind the impeller transition results from the strong flow cross-sectional taper and the associated static pressure build-up at this point. The reduction of the flow cross-section thus acts as a supersonic diffuser and prevents the formation of compression shocks. Although no compression shocks can be detected in this area, the total pressure loss Ap _dead is higher here than in the variant with a constriction x _min at a diameter of 300 mm (cf. 11 ). Accordingly, the short delay with the flow separation and the subsequent acceleration are more lossy than the compression shocks occurring in this region in the variant with a constriction x _min at a diameter of 300 mm.

Also, the supersonic free jet similar area shows in the variant according to the 10 a higher loss. This is because the variant according to the 11 in the length of the first section 48 sufficient to delay the gas flow in the narrowest cross-section (ie in the constriction x _min ) to subsonic. If the total pressure loss Ap _{dead of} both concepts over the entire (radial) length of the diffuser space 40 considered, both courses are similar. They are approximately linear. It should be noted, however, that from a radius of about 125 mm, the increase in the total pressure p _{tot is} steeper. This is due to the strong detachment that the gas flow experiences from here.

Due to the fact that the variant according to the 11 a lower total pressure loss Δp _Tot or a higher efficiency η _{D, s} shows, this was used for further optimization. To achieve this goal was a reduction of the flow cross section in the transition between the first section 48 and the second section 50 necessary. That showed preliminary investigations. Against this background, the flow cross-section of the variant of the diffuser space 40 according to the 11 in the transition between the two sections 48 . 50 at the diameter of 300 mm successively up to a width of 2.2 mm according to 12 reduced. As a result, the efficiency η _{D, s was increased} to 62% and the total pressure loss Ap _dead reduced to 1.27 bar (further). The flow changed in such a way that compared to 11 only the impact induced by the impeller transition can be recognized. The subsequent impact reflections hardly emerge and the expansion and compression areas of the free jet are no longer visible. This flow behavior is the result of the strong geometric flow cross-sectional taper, which acts as a supersonic diffuser.

Compared to the 11 is also in the 12 a deflection of the gas flow toward the diffuser ceiling from the transition between the two sections 48 . 50 to recognize. This is due to the replacement of the gas flow behind the impeller transition according to the 12 whereby the gas flow is deflected in the direction of constriction x _min to the diffuser floor and as a result undergoes a deflection in the direction of the diffuser ceiling in the region of the constriction x _min .

In addition, in the 12 to recognize that the core of the gas flow in contrast to the variant according to the 11 , That is accelerated again with a width of 3.2 mm at the narrowed x _min, from a radius of about 140 mm. The reason for this is that in this diffuser region, the flow separation decreases and the resulting flow cross-sectional widening for the flow core acts as a supersonic nozzle. From the constriction x _min the flow core retards, because the increasing flow separation has the effect of a cross-sectional taper. A further reduction in cross-section at the constriction would cause the gas flow in the direction of constriction x _min to be accelerated even more due to the aforementioned flow conditions, as a result of which the distance to the outlet is insufficient for a delay to subsonic noise. The specified for the diffuser exit subsonic condition would thus no longer satisfied. Further optimization by reducing the flow cross section at this point is therefore not possible.

In contrast to the embodiment of a diffuser space according to the invention 40 according to the 5 leads an embodiment according to the 6 to achieve the highest efficiency η _{D, s} at a position of the constriction x _min at a diameter of 220 mm (see. 13 to 15 ). The reason for this contradictory behavior lies in the different geometric forms of both embodiments. While the embodiment according to the 5 from the constriction x _{min has} a flow cross-sectional widening, the width of the diffuser space remains 40 (ie the distance between the diffuser space 40 delimiting housing walls) in the embodiment according to 6 constant. It follows that the embodiment according to the 6 a less strong flow separation in the free jet similar area experiences than the embodiment according to the 5 (see. 10 and 13 ). Despite this fact, the embodiment according to the 6 in the 13 and 14 a comparable flow behavior as the embodiment according to the 5 in the 10 and 11 , If the constriction x _{min is} at a diameter of 300 mm, compression impulses are visible behind the impeller transition, but not at a diameter of 220 mm. The flow areas which resemble a supersonic jet also occur with the difference that in the embodiment according to FIG 6 with the constriction x _min at a diameter of 220 mm more in the direction of diffuser exit is shifted. The reason for this lies in the constant width of the diffuser space 40 in the embodiment according to the 6 behind the constriction x _min , whereby the flow cross-section compared to the embodiment according to the 5 increases less over the radius. It follows that the flow cross-section, at which the gas flow separates, behind that in the embodiment according to the 5 although the given physical inlet and outlet conditions are the same.

Like the flow behavior, the course of the total pressure p _Tot in an embodiment of a diffuser according to the invention in accordance with FIG 6 with that in an embodiment according to the 5 comparable. The loss increase is approximately linear and increases more with the start of the free jet.

Despite these similarities, an optimization is equivalent to the embodiment according to the 5 for which the necking position with the highest efficiency η _{D, s was} selected, in the embodiment according to the 6 not possible. A further reduction in the width of the diffuser space 40 with the diameter of 220 mm, namely, would lead to a more pronounced supersonic nozzle shape and accelerate the gas flow more. The route length would no longer be sufficient for a delay of the flow and thus no longer fulfills the subsonic condition specified for the diffuser exit.

Against this background, the variant according to the 14 Ie x _min at a diameter of 300 mm, selected with the constriction and the flow cross section up to a width of 2.3 mm reduced (see FIG. 15 ). As a result, the efficiency η _{D, s was increased} to 69% and the total pressure loss Ap _dead reduced to 1.1 bar.

Likewise simulated variants of diffuser spaces according to the invention with necking positions between 220 mm and 300 mm showed a lower optimization potential due to the problem mentioned.

Compared to the variant according to the 12 Has reduction of a flow cross-section at constriction x _min at a variant in accordance with 15 similar effects on gas flow. The impact reflections are barely recognizable and the expansion and compression areas of the free jet no longer occur. In contrast to the variant according to the 12 the gas flow is deflected in the area of the constriction x _min but not in the direction of the diffuser ceiling. The reason for this is the constant width of the diffuser after the constriction x _min , which counteracts a deflection of the flow at this point. Another difference compared to the embodiment according to the 5 is that the acceleration of the flow core hardly occurs in the region of the constriction x _min . This is because the flow area at the constriction x _min with a width of 2.3 mm is smaller. A further reduction of the width at this point is still not possible, since the same effect as in the embodiment according to the 5 occurs. The gas flow would namely x _min accelerate in the direction narrowed more strongly, whereby the length of the route is not sufficient to the diffuser outlet for a delay to subsonic speed. The specified for the diffuser exit subsonic condition would thus no longer satisfied.

In summary, it should be noted that the embodiments according to the invention of diffusers according to the 5 and 6 to delay a gas flow loss less than a comparatively simulated parallel-walled diffuser, as in the US 3,010,642 is disclosed. In comparison, the variant improves according to the 12 the efficiency η _{D, s} 11% and reduces the total pressure loss .DELTA.p _Dead by 27%. The variant according to the 15 increases the efficiency η _{D, s} by 23% and reduces the total pressure loss Ap _dead by 37%. However, it should be noted that the surface-averaged exit velocity at the diffuser exit in accordance with 15 with 341 m / s (Mach = 0.87) above that of the variant according to the 12 with 266 m / s (Mach 0.67). The variant according to the 12 reaches the speed range of the parallel-walled comparison diffuser in terms of the area-averaged exit velocity; the variant according to 15 lies above it. With regard to the overall efficiency η _{D, s of} the compressor, these speed differences should be taken into account. The impingement of the spiral 38 namely, with a higher flow rate can cause higher losses in the spiral 38 to lead.

In addition, the calculations show that the static efficiency η _{D, s} and the total pressure loss Δp _Tot are equivalent as judgment criteria. If, for example, a diffuser shape shows a higher efficiency η _{D, s} , its total pressure loss Δp _Tot also drops.

LIST OF REFERENCE NUMBERS

10

 internal combustion engine

12

 cylinder

14

 crankshaft

16

 centrifugal compressors

18

 Intercooler

20

 throttle

22

 turbine

24

 wave

26

 wastegate

28

 belt drive

30

 compressor bypass

32

 Housing of the centrifugal compressor

34

 (Supersonic) compressor impeller

36

 inlet

38

 spiral

40

 diffuser space

42

 impeller channel

44

 impeller blade

46

 axis of rotation

48

 first section of the diffuser room

50

 second section of the diffuser room

c _m

 Meridian component of the absolute velocity of the gas flow

n

 Speed of the compressor impeller

A

 Flow area

m '

 mass flow

ρ

 density

v

 speed

A _critical

 critical flow cross-section

π _critical

 critical pressure ratio

u ₁

 Circumferential speed at the impeller inlet

u ₂

 Peripheral speed at impeller outlet

c ₁

 Absolute speed at the impeller inlet

w ₁

 Relative speed at the impeller inlet

c ₂

 Absolute speed at impeller outlet

w ₂

 Relative speed at impeller outlet

x _min

 constriction

p _dead

 Total pressure

Δp _dead

 Total pressure loss

η _{D, s}

 isentropic diffuser efficiency

QUOTES INCLUDE IN THE DESCRIPTION

This list of the documents listed by the applicant has been generated automatically and is included solely for the better information of the reader. The list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions.

Cited patent literature

• US 3010642 [0010, 0011, 0011, 0019, 0022, 0054, 0062, 0063, 0083]
• DE 1053714 [0012]
• US 3604818 [0013]
• JP 3260399A [0014]
• DE 1628227A [0015]

Claims (9)

Centrifugal compressor ( 16 ) with a housing ( 32 ) and a rotatable within the housing ( 32 ) mounted supersonic compressor impeller ( 34 ), the housing ( 32 ) an inlet ( 36 ) and an outlet, which are arranged such that the supersonic compressor impeller ( 34 ) was flowed by a gas to be compressed in the axial direction and flowed away in the radial direction, and wherein between the supersonic compressor impeller ( 34 ) and the outlet an annular, bounded by opposite housing walls diffuser space ( 40 ), characterized in that the housing walls in a first, of the supersonic compressor impeller ( 34 ) outgoing section ( 48 ) of the diffuser space ( 40 ) and, in a second, to the first section ( 48 ) subsequent section ( 50 ) run parallel to each other or move away from each other. Centrifugal compressor ( 16 ) according to claim 1, characterized in that the approach of the housing walls in the first section ( 48 ) and / or a removal of the housing walls from one another in the second section ( 50 ) is continuous. Centrifugal compressor ( 16 ) according to claim 2, characterized in that the approach of the housing walls in the first section ( 48 ) and / or a removal of the housing walls from one another in the second section ( 50 ) is linear. Centrifugal compressor ( 16 ) according to one of the preceding claims, characterized in that the flow cross-section of the diffuser space ( 40 ) above the radial height of the first section ( 48 ) reduced. Centrifugal compressor ( 16 ) according to one of the preceding claims, characterized in that the approach of the housing walls in the first section ( 48 ) and / or a removal of the housing walls from one another in the second section ( 50 ) is formed on both sides. Centrifugal compressor ( 16 ) according to one of the preceding claims, characterized in that the transition between the first section ( 48 ) and the second section ( 50 ) at least half the radial height of the diffuser space ( 40 ) is arranged. Centrifugal compressor ( 16 ) according to one of the preceding claims, characterized in that the transition between the first section ( 48 ) and the second section ( 50 ) is angled. Centrifugal compressor ( 16 ) according to one of the preceding claims, characterized in that the flow cross-section of one between two impeller blades ( 44 ) trained impeller channel ( 42 ) in at least a portion above the radial height of the impeller channel ( 42 ). Centrifugal compressor ( 16 ) according to one of the preceding claims, characterized in that the diffuser space ( 40 ) is unopened.

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