Complete engine model results

After calibrating the DIPulse combustion model through the CPOA, the model was implemented in the full engine model, replacing the previously used non- predictive combustion model, which employs experimentally measured burn rates. In the following figures (Performance in Figure 2.38, Flow in Figure 2.39, Pressure in Figure 2.40, Temperature in Figure 2.41) the comparison between the experimental data and the simulation results are reported for the full engine model. It should be noticed that the comparison was extended to more than 300 engine operating points, over the entire engine operating map.

Figure 2.38 – Experimental and simulated BMEP (left) and BSFC (right) comparison for the entire engine map

• BMEP (Figure 2.38 – left) – Since a controller was put in place to meet the BMEP target controlling the injected quantity by adjusting the duration of the main injection, the simulated BMEP always matches the measured value.

• BSFC (Figure 2.38 – right) – The model predicts with good accuracy the fuel consumption showing an average error lower than 4%. Significant errors are present at low load engine operating conditions (below 1 bar BMEP) which are characterized by experimental high dispersion. For these conditions, a very small error in FMEP gives high error in BSFC.

Figure 2.39 – Experimental and simulated volumetric efficiency comparison for the entire engine map

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• Volumetric efficiency (Figure 2.39) – Significant discrepancies between the simulated and the experimental results can be noticed. However, since pressure and temperature levels in the intake manifold are predicted with good accuracy, as it will be shown in the following section, the experimental mass flow rates do not appear to be always consistent with the other measurements.

• EGR fraction – since the EGR target map was not provided, the setpoints for the whole engine map had to be extrapolated from a limited dataset coming from other measurements.

Figure 2.40 – Experimental and simulated compressor outlet (left) and turbine inlet (right) pressure comparison for the entire engine map

• Compressor Outlet Pressure (Figure 2.40 – left) – In the simulation model the compressor outlet pressure is controlled by the acting on the VGT controller: therefore the simulated Compressor Outlet pressure should always match the corresponding experimental value.

• Turbine Inlet Pressure (Figure 2.40 – right) – The model predicts the turbine inlet pressure with very good accuracy.

Figure 2.41 – Experimental and simulated compressor outlet (left) and turbine inlet (right) temperature comparison for the entire engine map

• Compressor Outlet Temperature (Figure 2.41 – left) – The compressor outlet temperature predictions probably suffer from some turbocharger map inaccuracies and is not well matched for all the operating points. However this should not affect the in-cylinder

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calculations since the temperature downstream of the intercooler is imposed to be equal to the experimental value.

• Turbine Inlet Temperature (Figure 2.41 – right) – Turbine inlet temperature is very well matched despite it is normally a very critical parameter to meet with a 1D-CFD code due to the complex heat loss process in the exhaust port, exhaust manifold and turbine volute. A more detailed analysis of the DIPulse results can be found in Appendix A1 and A2. Appendix A1 is focused on the effects of the DIPulse calibration parameters on the combustion results: the analysis was carried out considering a variations of ±10% and ±30% (with respect to the optimized values) of each parameters while keeping all the other parameters constant. Appendix A2 reports the DIPulse results considering a sensitivity analysis on engine control parameters (as EGR rate, SOI, or boost pressure) as highlighted in Figure 2.33.

After the calibration and the validation of the detailed engine model in steady state conditions, it was reduced to a Fast Running Model (FRM) in order to be used for transient analysis, as the simulation of typical type approval driving cycles. The developed FRM model operates with a computational time 3 times slower with respect to real time and consists of:

• ECU controls in order to change engine calibration during the transient simulation, for example as a function of cooling water temperature or valve actuations.

• Thermal masses of pistons, liners, cylinder head and turbine for taking into account the thermal inertia in cold start conditions.

• Estimated friction losses based on Chen Flynn model as a function of cooling water temperature, as shown in Figure 2.42.

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The comparison between the experimental and simulated normalized fuel consumption along the WLTC is shown in Figure 2.43. The FRM 1D model underestimates the total fuel consumption of about 4%. This difference is mainly due to the difference in the first part of the driving cycle and the estimated Chen Flynn model at low coolant temperature may introduce errors in the prediction of real behavior. Nevertheless, the accuracy of the 1D model remains satisfactory, and the results of the simulated WLTC was taken as reference for the next analysis.

Figure 2.43 – Comparison between experimental (black dashed) and simulated (red solid) normalized fuel consumption over WLTC

As already stated, the accuracy with which the 1D engine model coupled with the injector model predicts the actual behavior of the system is more than satisfactory. In other words, a kind of virtual test rig was built in GT-SUITE, thanks to which the air management and fuel injection systems optimization, presented in the next chapters, was performed.

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Chapter 3

3

I part: Air management system

analysis

Part of the work described in this Chapter was also previously published in the following publications.

• Piano, A., Millo, F., Di Nunno, D., and Gallone, A., " Numerical Assessment of the CO2 Reduction Potential of Variable Valve Actuation on a Light Duty Diesel Engine," SAE Technical Paper 2018-37-0006, 2018,

https://doi.org/10.4271/2018-37-0006.

• Piano, A., Millo, F., Di Nunno, D., and Gallone, A., "Numerical Analysis on the Potential of Different Variable Valve Actuation Strategies on a Light Duty Diesel Engine for Improving Exhaust System Warm Up," SAE Technical Paper 2017-24-0024, 2017, https://doi.org/10.4271/2017-24- 0024.

As already stated in the Introduction, the future generation of Diesel engines must achieve 2 different goals: on one hand, there is the increasing need of efficiency improvements; on the other hand, the more stringent emissions regulations push the car manufacturer to produce low emissions vehicles. In this very complex context, an advanced and fully flexible air management system is becoming highly desirable for modern Diesel engines since it can help the engine development with the aims of reduce CO2 emission and improve the exhaust aftertreatment efficiency.

Differently from gasoline engines, where Variable Valve Actuation (VVA) is a well-established technique largely adopted in series production, variable air management systems in Diesel engines have not achieved considerable market penetration. However, their benefits are summarized as follows.

• Low-speed torque improvement retarding Exhaust Valve Opening (EVO) timing as suggested by Tai et al. [54]. Retarded EVO after

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Bottom Dead Center (BDC) results in an increase in turbine power (higher pressure pulses) and, consequently, in an air-to-fuel ratio increment, although with a significant Brake Specific Fuel Consumption (BSFC) penalty in steady state condition because of higher pumping losses and residuals within the cylinder. However, the BSFC penalty due to this technique can be negligible in the overall driving-cycle fuel consumption, due to the very short time spent at full load at very low engine speed. Same effect can be achieved by advancing EVO [55].

• Fuel consumption reduction. Engine efficiency improvement can be achieved by optimizing EVO timing at each engine speed, or enabling Miller cycle obtaining higher thermodynamic efficiency with a very high geometric compression ratio. [25,56–59]

• Emission reduction. Theoretically, adjusting IVC with a flexible valve actuation leads to an air–fuel ratio increase and a consequent soot reduction. Another effect is related to the reduced effective compression ratio by early or late IVC to obtain lower compression pressure, combustion temperature, and NOx. [60–63] Negative valve overlap, intake valve pre-opening during the intake stoke, or exhaust valve post- opening during the intake stroke, could enable internal EGR, as effective way for reducing engine-out emissions. [62,64]

• Turbulence modulation within the cylinder. Gas motion level can be adjusted by using unequal intake valve lifts thus enhancing the swirl motion in order to reduce soot formation at low speed). [56] Thanks to the adoption of VVA, the redundant air control valve (intake and exhaust throttle) could be eliminated, simplifying the overall engine system. [56]

• Advanced combustion modes enabling, such as Homogeneous Charge Compression Ignition (HCCI). VVA can provide variable effective compression ratio, charge composition and temperature control required for such advanced combustion modes. [65–67]

• Aftertreatment system performance improvement. An additional opening of the exhaust valve during the intake stroke, or an early EVO could provide higher exhaust temperature that can speed-up the warm up of the aftertreatment system since to the higher in-cylinder residuals [68–70].