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New Approach for Centrifugal Compressor Calculations

Profile image of Ronny WernerRonny Werner

2013, MTZ worldwide

Abstract

Exhaust gas turbocharging has become established as a decisive tool in the implementation of downsizing concepts. Further optimisation requires reliable simulation of the processes taking place in the turbocharger and its radial compressor. At the Chair of Internal Combustion Engines at the TU Dresden, a new concept for the CFD simulation of a radial compressor and its interaction with engine process simulation has been developed. The basic approach is the implementation of variable throttling within the compressor system. As a result, the operating behaviour can be described during the calculation independently of the previous measurement results. 10 COVER STORY SUPERCHARGING ❽ Swirl number in the near-wall region at constant turbocharger speed and varying mass flow

Figures (12)
Explanation of symbols used
Explanation of symbols used
According to the requirements of inde- pendence of experimentally determined specifications, the calculation of the investigated CFD model using pressure boundary conditions is carried out both at the inlet and at the outlet. The maxi- mum throughput for the respective com- pressor speed results directly from a sim- ulation with fully continuous permeabil- The experimental investigations were carried out at the new hot gas test bench
According to the requirements of inde- pendence of experimentally determined specifications, the calculation of the investigated CFD model using pressure boundary conditions is carried out both at the inlet and at the outlet. The maxi- mum throughput for the respective com- pressor speed results directly from a sim- ulation with fully continuous permeabil- The experimental investigations were carried out at the new hot gas test bench
With the help of the developed simula- tion model, operating points below the resistance performance curve of the test bench i.e. inverse pressure conditions with p,/p, < 1, can also be represented. By changing the boundary conditions each compressor performance curve can be extended up to the fourth quadrant. at the Chair of Combustion Engine Fac- ulty of the TU Dresden, @. The plant has two separate combustion chambers with 200 and 800 kW of power respectively, in which both passenger car and commer- cial vehicle engine turbochargers can be measured. The level of compression can be increased by means of the operation of a closed compressor circuit (closed loop). As a result of the increased com- pression performance, the measurable
With the help of the developed simula- tion model, operating points below the resistance performance curve of the test bench i.e. inverse pressure conditions with p,/p, < 1, can also be represented. By changing the boundary conditions each compressor performance curve can be extended up to the fourth quadrant. at the Chair of Combustion Engine Fac- ulty of the TU Dresden, @. The plant has two separate combustion chambers with 200 and 800 kW of power respectively, in which both passenger car and commer- cial vehicle engine turbochargers can be measured. The level of compression can be increased by means of the operation of a closed compressor circuit (closed loop). As a result of the increased com- pression performance, the measurable
Hot gas test bench at the chair of combustion engines faculty operating range of the turbine can be significantly increased. This extended turbine measurement produces an addi- tional data base, which makes possible a more accurate extrapolation of the data for the 11) simulation nrocress.
Hot gas test bench at the chair of combustion engines faculty operating range of the turbine can be significantly increased. This extended turbine measurement produces an addi- tional data base, which makes possible a more accurate extrapolation of the data for the 11) simulation nrocress.
6) Efficiency comparison between simulation and measurement
6) Efficiency comparison between simulation and measurement
@ Flow behaviour in the centrifugal compressor at different mass flow rates and constant turbo speed The efficiency ratio comparison, ©, clearly reveals the effect of the heat flow from the turbine into the compressor. Even at higher speed a deviation of 2 % between the measured and calculated efficiency is apparent. In the case of lower speeds this effect is much more pronounced, so that the difference at 165 m/s is increased by 20 % points. The causes are to be sought in the lack of heat transfer from the turbine to the com- pressor compared to the measurements. Furthermore, it is assumed in the simula- Another advantage of the simulation is the possibility to extend the compressor performance at low peripheral speeds. In contrast to conventional turbocharger measurements all operating conditions between zero and maximum speed are displayed. With regard to the engine pro- cess simulation an extended database can
@ Flow behaviour in the centrifugal compressor at different mass flow rates and constant turbo speed The efficiency ratio comparison, ©, clearly reveals the effect of the heat flow from the turbine into the compressor. Even at higher speed a deviation of 2 % between the measured and calculated efficiency is apparent. In the case of lower speeds this effect is much more pronounced, so that the difference at 165 m/s is increased by 20 % points. The causes are to be sought in the lack of heat transfer from the turbine to the com- pressor compared to the measurements. Furthermore, it is assumed in the simula- Another advantage of the simulation is the possibility to extend the compressor performance at low peripheral speeds. In contrast to conventional turbocharger measurements all operating conditions between zero and maximum speed are displayed. With regard to the engine pro- cess simulation an extended database can
© Swirl number in the near-wall region at constant turbocharger speed and varying mass flow @ shows the flow processes in the area of the inflow path as well as the com- pressor wheel for a constant compressor peripheral speed of 400 m/s at various operating points. In the area of the surge line the formation of a swirling flow upstream of the compressor can be clearly seen. Within the steady-state analysis the phenomenon of pumping, contrary to the real behaviour, cannot be
© Swirl number in the near-wall region at constant turbocharger speed and varying mass flow @ shows the flow processes in the area of the inflow path as well as the com- pressor wheel for a constant compressor peripheral speed of 400 m/s at various operating points. In the area of the surge line the formation of a swirling flow upstream of the compressor can be clearly seen. Within the steady-state analysis the phenomenon of pumping, contrary to the real behaviour, cannot be
The swirl upstream of the compressor wheel, which results from the backflow, is used for the determination of the surge line in the CFD calculation. The basis here is the definition of a swirl number according to Eq. 4, which arises from the rotational and axial momentum, Eq. 5 and Eq. 6.Here, u and w are defined as the axial and tangential velocity compo- nent respectively.
The swirl upstream of the compressor wheel, which results from the backflow, is used for the determination of the surge line in the CFD calculation. The basis here is the definition of a swirl number according to Eq. 4, which arises from the rotational and axial momentum, Eq. 5 and Eq. 6.Here, u and w are defined as the axial and tangential velocity compo- nent respectively.
© Simulated engine characteristics at full load for different data sets of the compressor characteristic map in comparison to the measurement Significantly greater differences are recorded by the evaluation of the load steps. Examples of this course, the selected specifications for an engine speed of nyo = 2000 rpm are shown over time in @. An initial operating point with a mean effective pressure of 2 bar results in a spontaneous pedal- travel value increase at 0 s of up to 100 %. Using the results obtained, the torque build-up of the engine and the In @ a comparison between the calcu- lated engine performance data and the test bench results (reference) is illus- trated for the full-load operation of a 2.0- diesel engine. The simulation was backed up, in the first case, by the meas- ured compressor performance map of the hot-gas test bench, and in the second case, by the data of the simulated com- pressor performance map in the 1D model. In the area of the lower engine speeds, the results show that, by use of Already at the start point before the load step, the differences in specific fuel consumption are clearly evidenced. Thus, in the measurement on the test bench, a value of 258 g/kWh was calcu- lated in the simulation with measured compressor performance map 273 g/kKWh (+6 %) and with the CFD compressor performance map 261 g/kWh (+1.1 %). During the load step, significant differ- ences occur between the values of the engine test and the simulation results with experimental data base. Owing to an underestimated compressor efficiency ratio based on performance map meas- urements, a delayed compressor speed- up occurs, which has an adverse effect on the charge and the mean pressure build-up (+0.4 s). Using the results from the CFD calculation compressor perfor- mance map, these disadvantages can be minimised and so the readings of the engine test bench can be almost exactly reproduced.
© Simulated engine characteristics at full load for different data sets of the compressor characteristic map in comparison to the measurement Significantly greater differences are recorded by the evaluation of the load steps. Examples of this course, the selected specifications for an engine speed of nyo = 2000 rpm are shown over time in @. An initial operating point with a mean effective pressure of 2 bar results in a spontaneous pedal- travel value increase at 0 s of up to 100 %. Using the results obtained, the torque build-up of the engine and the In @ a comparison between the calcu- lated engine performance data and the test bench results (reference) is illus- trated for the full-load operation of a 2.0- diesel engine. The simulation was backed up, in the first case, by the meas- ured compressor performance map of the hot-gas test bench, and in the second case, by the data of the simulated com- pressor performance map in the 1D model. In the area of the lower engine speeds, the results show that, by use of Already at the start point before the load step, the differences in specific fuel consumption are clearly evidenced. Thus, in the measurement on the test bench, a value of 258 g/kWh was calcu- lated in the simulation with measured compressor performance map 273 g/kKWh (+6 %) and with the CFD compressor performance map 261 g/kWh (+1.1 %). During the load step, significant differ- ences occur between the values of the engine test and the simulation results with experimental data base. Owing to an underestimated compressor efficiency ratio based on performance map meas- urements, a delayed compressor speed- up occurs, which has an adverse effect on the charge and the mean pressure build-up (+0.4 s). Using the results from the CFD calculation compressor perfor- mance map, these disadvantages can be minimised and so the readings of the engine test bench can be almost exactly reproduced.
@® Simulated engine parameters for a load step at 2000 rpm for different data sets of the compressor characteristic map in comparison to the measurement
@® Simulated engine parameters for a load step at 2000 rpm for different data sets of the compressor characteristic map in comparison to the measurement

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