The invention relates to a method for operating a supercharged internal combustion engine.
German Patent Application No. DE 102 21 014 A1 describes an internal combustion engine with exhaust gas turbocharger, the exhaust gas turbine of which is equipped with a variable turbine geometry for variably adjusting the turbine inlet cross section. The variable turbine geometry is adjusted in such a manner that the exhaust gas turbocharger rotational speed of the exhaust gas turbocharger is within a predetermined, permissible rotational speed range. This is achieved by virtue of the fact that, for example at low engine speeds and engine loads, the variable turbine geometry is adjusted in the direction of its back-up position, in which the free turbine inlet cross section adopts a minimum, after which the exhaust gas back pressure between internal combustion engine and exhaust gas turbine rises and the exhaust gas is passed through the remaining free cross section of flow at a high velocity and strikes the turbine wheel at this high velocity. It is in this way possible to keep the exhaust gas turbocharger rotational speed at a desired minimum level.
Furthermore, German Patent Application No. DE 102 21 014 A1 discloses running the compressor in turbine mode provided that, at low loads and speeds of the internal combustion engine, a sub-atmospheric pressure is present in the intake section immediately ahead of the cylinder inlets, with the result that a pressure drop is produced across the compressor, which can be used to drive the compressor impeller. This operating mode is also known as cold-air turbine operation of the compressor. Moreover, the compressor is assigned an additional drive which is used to compensate for an energy deficit of the exhaust gas turbine at certain operating points of the internal combustion engine. The rotational speed of the exhaust gas turbocharger can be kept approximately constant with the aid of the additional drive.
In turbine mode of the compressor, energy is fed to the charger by actuating the additional drive, in order to increase the exhaust gas turbocharger rotational speed, with the result that the compressor can be operated in the region of its optimum efficiency during its cold-air turbine mode. In this case, however, it should be taken into account that the exhaust gas turbine arranged in the exhaust section, on account of the increase in the exhaust gas turbocharger rotational speed, passes into an efficiency range at which the turbine begins to ventilate and consumes power, which has a braking effect on the exhaust gas turbocharger rotational speed.
It is therefore an object of the invention to operate a supercharged internal combustion engine, the exhaust gas turbine of which is equipped with a variable turbine geometry, in such a manner as to produce efficiency-optimized operation of the exhaust gas turbocharger. In particular in the lower load/speed range of the internal combustion engine, in which the compressor is operated in cold-air turbine mode, both the compressor and the exhaust gas turbine should be operated in the region of their respective optimum efficiencies.
In the method according to the invention for operating a supercharged internal combustion engine, the exhaust gas turbine of the charger is equipped with a variable turbine geometry for variably adjusting the effective turbine inlet cross section between a minimum build-up position and a maximum open position. In the lower load/speed range of the internal combustion engine, in which the pressure upstream of the compressor is higher than the pressure downstream of the compressor and the compressor is operating in what is known as the cold-air turbine mode, the variable turbine geometry of the exhaust gas turbine is adjusted in the direction of its back-up position until the turbine efficiency of the exhaust gas turbine is at least approximately in the region of the optimum efficiency.
This defines a directly dependent relationship between the fast running speed of the exhaust gas turbocharger which is to be set and the narrowest turbine cross section, which is reached in the back-up position or at least close to the back-up position of the variable turbine geometry. It is in this way possible, in particular in cold-air turbine mode of the compressor, in which there is a pressure drop across the compressor, which is utilized to drive the compressor impeller, to operate both the compressor and the cold-air turbine in the region of their optimum efficiencies and also to operate the exhaust gas turbine in the region of its optimum efficiency. Increasing the turbine pressure ratio also increases the isentropic expansion rate and therefore also the turbine power of the exhaust gas turbine, which means that despite the higher rotational speed of the exhaust gas turbine, the optimum efficiency range is not departed from, and in particular an undesired ventilation mode, in which energy is consumed, is avoided.
In principle, these measures make it possible to dispense with an additional drive for the charger without the risk of efficiency losses or a drop in the charger rotational speed. Rather, the charger rotational speed is kept at an approximately constant and high level. Nevertheless, it may be expedient to provide an additional drive.
The compressor of the exhaust gas turbocharger expediently has an additional passage, which is formed separately from the compressor inlet passage and opens out radially into the compressor inlet passage at the compressor impeller. The combustion air stream which is to be supplied via the additional passage is adjustable, with the combustion air stream which is to be supplied being passed via the additional passage in particular in the lower load/speed range of the internal combustion engine, this air stream then striking the compressor impeller blades radially and imparting a driving momentum to them. Due to the pressure gradient across the compressor, combustion air is sucked in from the environment. The compressor which is operated in cold-air turbine mode makes a contribution to maintaining the charger rotational speed. As the load or speed of the internal combustion engine increases, it is possible to reduce the supply of air across the additional passage and ultimately to eliminate this supply of air altogether, so that the combustion air takes the normal path via the compressor inlet passage and strikes the compressor impeller at the end side. At higher loads and speeds of the internal combustion engine, the compressor is operated in compressor mode, with the combustion air which is supplied being compressed to an increased boost pressure.
To eliminate the risk of excessive rotational speeds in the rotor of the exhaust gas turbocharger, the variable turbine geometry, if the exhaust gas turbocharger rotational speed exceeds an upper limit value, can be adjusted in the direction of its open position until the incoming flow exerts a braking action on the turbine wheel, after which the exhaust gas turbine consumes energy and has a braking action on the exhaust gas turbocharger rotational speed. This operating mode is also known as ventilation mode of the exhaust gas turbine. The risk of excessive rotational speeds may occur in particular in the event of load changes in the internal combustion engine from a high load towards a low part-load, which is associated with a considerable pressure drop in the intake section immediately upstream of the cylinder inlets. As a result, the load on the compressor is greatly relieved and it suddenly shifts to cold-air turbine mode, in which the compressor delivers drive energy to the rotor. At the same time, the hot exhaust manifold is responsible for supplying considerable energy to the exhaust gas, with the result that the exhaust gas turbine is also briefly providing further drive energy, which overall would lead to an unacceptably high rise in the exhaust gas turbocharger rotational speed. To avoid this, the variable turbine geometry of the exhaust gas turbocharger is opened as quickly as possible to a sufficient extent for the efficiency of the exhaust gas turbine to become negative and the turbine to be operated in ventilation mode, in which energy is consumed. The resulting, negative power resulting from bearing friction and braking power of the exhaust gas turbine must be greater than the driving power of the compressor which is being operated in cold-air turbine mode.
After the braking influence of the exhaust gas turbine has reduced the exhaust gas turbocharger rotational speed to a permissible level, the variable turbine geometry can return to the position appropriate for the current operating mode, i.e. low loads and speeds of the internal combustion engine can in particular be moved back towards the back-up position.
Further advantages and expedient embodiments are given in the further claims, the description of the figures and the drawings, in which:
In the figures, identical components are provided with identical reference designations.
The internal combustion engine 1 illustrated in
The compressor 5 is equipped with a variable compressor geometry 10, by means of which one of a total of two flow cross sections via which combustion air can be fed to the compressor impeller can be variably adjusted. The cross section of flow with the variable compressor geometry 10 is located in an additional passage 11, which branches off from the compressor inlet passage upstream of the compressor impeller and opens back out into the compressor inlet passage in the radial direction at the compressor impeller. The combustion air stream which is supplied via the additional passage 11 is controlled with the aid of an adjustable blocking valve 12 which can be regulated between a position in which it blocks the additional passage 11 and a position in which it opens the additional passage 11, and moreover can also block or open the axial compressor inlet passage. At medium and high loads or speeds of the internal combustion engine, combustion air is fed onto the compressor impeller in the axial direction via the axial compressor inlet passage, and the compressor impeller is driven by the turbine wheel via the shaft 7 and compresses the combustion air which is supplied to an increased boost pressure. In this operating mode, the additional passage 11 is expediently blocked off. At low loads and speeds of the internal combustion engine, by contrast, the combustion air is supplied via the additional passage 11; in this operating phase, a sub-atmospheric pressure is present immediately upstream of the cylinder inlets, with the result that a pressure gradient is established across the compressor, which can be utilized to drive the charger. The combustion air which is supplied via the additional passage 11 in this operating phase strikes the compressor impeller blades in the radial direction and imparts a driving momentum to them. The flow path across the compressor inlet passage is expediently blocked off in this operating phase.
Upstream of compressor 5, there is an air filter 13 in the intake section, in which air filter the combustion air supplied is purified. An air gauge 14, which is used to measure the air throughput and feed this information as a signal to a control unit 21, is arranged in intake section 6, downstream of air filter 13 and upstream of compressor 5. A charge air cooler 15, in which the combustion air is cooled, is arranged in intake section 6 downstream of compressor 5. Following charge air cooler 15, the combustion air is fed to the cylinders of internal combustion engine 1.
On the exhaust gas side, the exhaust gases which have been expelled from the internal combustion engine are passed via exhaust section 4 to exhaust gas turbine 3, it being possible to influence the efficiency of the exhaust gas turbine by way of the current setting of the variable turbine geometry between back-up position and any desired intermediate positions all the way through to the open position. If appropriate, blow-off via bypass 8 may also be suitable, if blow-off valve 9 is open. Downstream of exhaust gas turbine 3, the exhaust gases are purified in an exhaust gas purification unit 16.
Furthermore, an exhaust gas recirculation device 17, which comprises a recirculation line 18 between exhaust section 4 upstream of exhaust gas turbine 3 and intake section 6 of charge air cooler 15, is provided in the internal combustion engine. In return line 18 there is a controllable recirculation blocking valve 19 and an exhaust gas cooler 20. With the aid of the exhaust gas recirculation, it is possible to produce the NOx emissions in particular in the part-load range of the internal combustion engine.
All the adjustable units of the internal combustion engine can be adjusted as a function of state and operating variables by means of a control unit 21. This applies in particular to blocking valve 12, which can be used to control the supply of air into the compressor inlet passage or additional passage 11, the position of variable compressor geometry 10, recirculation blocking valve 19 in recirculation line 18 of exhaust gas recirculation device, and the position of variable turbine geometry 22.
An additional passage 33, which likewise branches off from the upstream air collection space 26 and opens back into the compressor inlet passage radially via an opening region 34, with opening region 34 in axial terms located at the level of the compressor impeller blades 24 of compressor impeller 23, is provided parallel to compressor inlet passage 25 but offset radially outwards and separated from the compressor inlet passage by means of an axial slide 32. In the position of blocking member 28 illustrated in
Axial slide 32 is acted on by axial forces from a spring element 37 and is pushed away from fixed swirl grate 35. The result of this is that when blocking member 28 moves away from compressor impeller 23, axial slide 32 is also displaced in the same direction as the blocking member by the force of spring element 37, with the result that the free cross section of flow in opening region 34 is increased. The opening movement of axial slide 32 is delimited by a stop 39.
Compressor 5 adopts the position of the blocking member 28 illustrated in
As the load and speed increase, blocking member 28 is moved back until entry opening 27 of the compressor inlet passage 25 is opened. Then, combustion air can flow out of air collection space 26 via entry opening 27 into compressor inlet passage 25 and strike the compressor impeller 23 in the axial direction. In this operating mode, the compressor is performing compressor work and compresses the combustion air which is supplied to an increased boost pressure. On the other hand, if there is a pressure drop across the compressor, the combustion air which is supplied via the additional passage 33 delivers a momentum which drives the compressor impeller.
Variable turbine geometry 22 is illustrated in detail in
as a function of the heat capacity cp of the gas which is in each case flowing through, the temperature Tt,in at the compressor impeller or turbine wheel inlet, the compressor or turbine pressure ratio pt,in/pt,out and the isentropic exponent κ. According to this relationship, the isentropic expansion rate co rises as the pressure ratio pt,in/pt,out increases. For the exhaust gas turbine, this means that its isentropic expansion rate co increases if the variable turbine geometry is adjusted in the direction of the back-up position, since with a smaller turbine inlet cross section the exhaust gas back pressure and therefore also the pressure ratio across the exhaust gas turbine increases.
The circumferential velocity u is calculated from
u=DCAπnETC
u=DHTπnETC
for the compressor (cold air mode “CA”) or for the exhaust gas turbine (hot turbine “HT”) as a function of the diameter DCA of the compressor impeller or the diameter DHT of the turbine wheel, in each case measured on the blade leading edge, as shown in
The ratio u/c0 of circumferential velocity u and isentropic expansion rate co rises with an increasing charger rotational speed nETC and drops with an increasing exhaust gas back pressure, on account of the expansion rate c0 then rising. If the variable turbine geometry is adjusted in the back-up position, the exhaust gas back pressure increases and the ratio u/c0 decreases.
In both the curves shown in
To prevent unacceptably high rotational speeds of the rotor of the exhaust gas turbocharger, which can occur in particular in the event of load changes from a high load on the internal combustion engine toward a low part-load, since in this case a sub-atmospheric pressure is established on the compressor side downstream of the compressor, which additionally drives the compressor but at the same time supplies reheating energy from the exhaust manifold on the exhaust gas side for heating and a correspondingly high energy in the exhaust gas, with the result that drive energy is likewise additionally acting on the charger shaft, in this case the variable turbine geometry is adjusted in the direction of its open position until the efficiency curve shown in
After the exhaust gas turbocharger rotational speed has reached a desired rotational speed level, in particular has dropped back to a lower rotational speed level, the variable turbine geometry can be adjusted back in the direction of its back-up position, in order to move the exhaust gas turbine out of the ventilation mode, in which it consumes energy, into the driving mode, in which it outputs energy. This adjustment of the variable turbine geometry in the direction of its back-up position can be carried out as soon as a parameter assigned to the exhaust gas turbocharger rotational speed adopts a defined value, for example if the exhaust gas turbocharger rotational speed is below a defined limit value for a minimum period of time.
Accordingly, while only a few embodiments have been have been shown and described, it is obvious that many changes and modifications may be made thereunto without departing from the spirit and scope of the invention.
Number | Date | Country | Kind |
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102004051889.0 | Oct 2004 | DE | national |