Effect of Supercharging Pressure on Internal Combustion Engine Performances and Pollutants Emissions

Effect of supercharging pressure on internal combustion engine performances and pollutants emissions

P. Podevin ; G. Descombes C. Charpentier Conservatoire National des Arts et Métiers Université P. et M. Curie Paris 6 Chaire de turbomachines et moteurs Laboratoire de mécanique physique 292 rue Saint Martin CNRS UPRESA 7068 75141 Paris Cedex 03 France 2, place de la gare de ceinture Tel: 33 (0)1 30 45 87 35 78210 Saint-Cyr l’Ecole Fax: 33 (0)1 30 45 87 73 Tel: 33 (0)1 30 85 48 81 E-mail: [email protected]

ABSTRACT The first advantage of turbocharging is the increase of specific power. Automotive engine requires a high torque at low engine speed in order to improve acceleration of the vehicle that means to have a high inlet pressure. Nevertheless, this high pressure is difficult to get with conventional or variable geometry turbocharger, the effect of supercharging for low engine speed has been studied, in order to estimate the advantages or disadvantages obtained in terms of performances and pollution. The experimental study is led on a 2.1 liter turbocharged indirect injection engine. Performances and following gas emissions CO, HC, NOx, CO2, O2, smoke have been measured. The most significant results are presented in this paper. Results can lead to an evaluation of the benefits expected.

NOMENCLATURE AFR Air fuel ratio QLVH Lower heating value of fuel Bmep Brake mean effective pressure Rpm Rotation per minute Bsfc Brake specific fuel consumption SN Smoke number Bosch unit FID Flame Ionisation Detector Tem Exhaust manifold temperature °C HC Unburned Hydrocarbon φ Equivalence ratio: AFRstochio/AFR MPA Magnetopneumatic Analyser ηg Gross efficiency N Engine speed in rpm ηv Volumetric efficiency NDIR Non-dispersive Infrared Analyser ρie Density of air at engine inlet NOx Nytrogen oxydes Pie Relative pressure at engine inlet

2 INTRODUCTION It is usually considered that turbocharging leads to fuel consumption reduction and diminution of pollutant emissions. This argument is not applicable to the whole area of engine functioning. For the automotive turbocharged Diesel engine, it is required a high torque at low engine rotation speeds. Hence, the turbine wheel is oversized in order to give a noticeable power to the compressor in this range. So, this arrangement leads to extreme exhaust counter- pressure at high loads and high speeds, whatever the means used for the control of the boost pressure (waste-gate (Anada et al 1997 [1]), variable scroll (Toussaint et al 1999 [2]), adjustable guide-vanes (Descombes et al 1999 [3)]). In spite of this oversizing, the available power is at very low speed insufficient to give a substantial boost pressure. Even if the means to obtain this substantial boost pressure are difficult to achieve, it is of a major interest to estimate its effect on engine performances to appreciate the potential benefits. The study of the effect of boost pressure can’t be done on a usual turbocharged engine due to the link between the compressor and the turbine and more generally between the compressor, the engine and the turbine. So, to overcome this difficulty, the engine is fed with compressed air at different pressure levels. This experimentation has been done on low engine speed. Most significant parameters in terms of performances and pollutant emissions have been registered.

1. EXPERIMENTAL SETUP Testes are performed on a standard test bench, fitted with a flywheel in order to simulate vehicle acceleration (Podevin et al 1999 [4]). The general layout of the test rig is presented in Figure 1. In order to get an adjustable boost pressure, the following alterations have been done to the engine: - only the standard turbine scroll has been fitted to the exhaust manifold. The remaining hole has been obstructed. -an additional turbocharger has been set close to the engine. The turbine is supplied with dry compressed air controlled by a pneumatic air regulator. This design has been retained as we have at our disposal a large amount of compressed air (700 m3 at 25 bar). The benefit of such arrangement is that the original setting and functioning of the engine is nearly kept. Furthermore, the use of this additional turbocharger allows to set the boost pressure in a very accurate and easy way. The main characteristics of the engine are: - Turbocharged Diesel - 4 cylinders - 12 valves - Displacement: 2088 cm3 - Compression ratio: 21.5: 1 - Power: 81 kW at 4300 rpm - Peak torque: 248 N.m at 2000 rpm 3 15 14 12 13 12 5 10

2

11 4 3 6 7 1

16 9 8

1 POSITIVE DISPLACEMENT METER 5 HEAT EXCHANGER AIR/WATER 9 EXHAUST 13 FLY WHEEL 2 TRANQUILLIZATION TANK 6 ENGINE 10 COMPRESSED AIR 14 DYNAMOMETER 3 AIR FILTER 7 TURBINE CASING 11 CARDAN JOINT 15 ELECTROMAGNETIC CLUTCH 4 ADDITIONAL TURBOCHARGER 8 MUFFLER 12 ELASTIC COUPLING 16 PNEUMATIC ENGINE

Fig. 1 - Sketch of the test bench

The engine was fully instrumented for temperature and pressure measurements, including in pre-chamber pressure. Gas emissions capability included: - CO2, CO, NOx (NDIR analyzer) - HC (FID analyzer) - O2 (MPA analyzer) smoke measurements: - Smoke Number (AVL 409 Smoke Meter) - In line Opacity Meter (Celesco 107)

2. TEST CONDITIONS The main interest of this study concerns the potential benefit we can obtain with an increase of boost pressure at low engine speeds. So, this research has been done for speeds 1500, 2000, 2500 rpm with different pressure rates:

Engine speed Inlet engine pressure (rpm) (mbar) 1500 0, 200, 400, 600 2000 0, 300, 600, 900 2500 0, 250, 500, 750, 1000

For these experiments, the exhaust gas recirculation device EGR was disconnected. No adjustments have been done to the injection system.

3. RESULTS 3.1 Influence of boost pressure on fuel consumption In Figure 2, is represented the brake specific fuel consumption Bsfc versus torque at 1500 rpm for different inlet pressure. When the inlet pressure rise, the benefits in torque and 4 Bsfc can be clearly stated. Increase of torque seems logical because more energy (more fuel) can be put when the boost pressure rises. Improvement in Bsfc can be demonstrated. For this purpose, it seems useful to express the engine performances as a function of brake mean effective pressure Bmep, gross efficiency ηg and the equivalence ratio φ (Figure 3).

Bsfc (g/kw.h) Gross efficiency (%) Bmep (mbar) 500 15000

Pie = 0 14000 Pie = 200 mbar Pie = 400 mbar 40 13000 450 Pie = 600 mbar 12000

35 11000 400 10000

30 9000

350 8000

25 7000

0

= 6000

r r r 300 a a a b Pie b b m m m 20 5000 0 0 0 0 0 0 2 6 4 = = = 4000 ie e ie P i 250 P P 15 3000

2000 Torque (N.m) Equivalence ratio 200 10 1000 0 40 80 120 160 200 240 280 0.00.20.40.60.81.01.2 Fig. 2 – Bsfc versus torque for different relative engine Fig. 3 – Bmep and ηg versus equivalence ratio φ for inlet pressure different relative engine inlet pressure at N = 1500 rpm at N = 1500 rpm

The brake mean effective pressure can be expressed by the following equation (Magnet 1998 [5]):

η v ρie ηg Q LVH Eq. Bmep = (1) AFR For example, we will compare the brake mean effective pressure at equivalence ratio 0.8 for point A with relative boost pressure 0 and point B with 600 mbar.

Bmep ρie B ηv B ηg Eq. B = B (2) Bmep ρ η η A ie A v A g A

Bsfc and ηg can be read in Figure 3, extra experimental value necessary for this calculation are given below:

Point Bmep ρie ηv ηg bar A 7.1 1.192 0.85 0.31 B 12.5 1.855 0.9 0.33

Bmep So, we have: B =1.556 ⋅1.059 ⋅1.065 =1.754 which is in accordance with BmepA Bmep 12.5 the ratio B = =1.76 BmepA 7.1 5 This result shows that increase of brake mean effective pressure is not only due to the rise of air density but also to the improvement of volumetric efficiency and gross efficiency. The rise of these two last terms is also the reason of lower brake specific fuel consumption. This result is logical: - for high inlet pressure there is a better filling of the cylinder, so a better volumetric efficiency, - increase of gross efficiency is due to the fact that a certain amount of power is given by compressed air. This point will be discussed later on. In Figure 4, same kinds of results are shown for 2500 rpm engine rotation speed.

Gross efficiency (%) Bmep (mbar) 15000 Pie = 0 Pie = 250 mbar 14000 Pie = 500 mbar Pie = 750 mbar 40 Pie = 1000 mbar 13000 12000

35 11000

10000

30 9000

8000

25 7000

6000

20 5000

4000

15 3000

2000 Equivalence ratio 10 1000 0.00.20.40.60.81.01.2

Fig. 4 – Bmep and ηg versus equivalence ratio φ for different relative engine inlet pressure at N = 2500 rpm

3.2 Influence of boost pressure on smoke emissions As for the last curves, we choice to represent smoke emissions versus equivalence ratio. Results are presented in Figures 5 and 6 respectively for 1500 and 2500 rpm engine speed.

Smoke number (Bosch Unit) Opacity (%) Smoke number (Bosh unit) Opacity (%) 8 40 8 40 Pie = 0 Pie = 200 mbar Pie = 0 7 Pie = 400 mbar 7 Pie = 250 mbar Pie = 600 mbar Pie = 500 mbar 6 30 6 Pie = 750 mbar 30 Pie = 1000 mbar 5 5

4 20 4 20

3 3 Smoke number Smoke number 2 10 2 10 Opacity Opacity 1 1 Equivalence ratio Equivalence ratio 0 0 0 0 0.0 0.2 0.4 0.6 0.8 1.0 1.2 0.00.20.40.60.81.01.2 Fig. 5 – Smoke number and opacity versus equivalence Fig. 6 – Smoke number and opacity versus equivalence ratio φ for different relative engine inlet pressure ratio φ for different relative engine inlet pressure at N at N = 1500 rpm = 2500 rpm

6 Both smoke number and opacity curves have been drawn even they are representatives of the same phenomena. At 1500 rpm, it can be noticed that smoke emissions are close whichever the inlet engine pressure. At 2500 rpm, some differences appear for equivalence ratio greater than 0.75 . Equivalence ratio 0.75 can be considered for a diesel engine as maximum admissible value. Greater values lead to a quick increase of black smoke emission. So, if smoke emissions is consider as representative of combustion effectiveness, results show that combustion depends particularly on the equivalence ratio rather than on the boost pressure.

3.3 Influence of boost pressure on gas emissions concentration 3.3.1 Nytrogen Oxyde: NOx NOx depends mostly on equivalence ratio and on maximum temperature in combustion chamber. In consequence, NOx emission is indirectly linked with manifold exhaust temperature. For engine speed 1500 rpm (Figure 7), the variation of NOx versus equivalence ratio is about the same for the different boost pressure. At 2000 rpm there is a slight difference. A higher production of NOx when the boost pressure increase (Figure 8). The same phenomenon appears at 2500 rpm but this tendency is not so obvious.

NOx (ppm) Tem (°C) NOx (ppm) Tem (°C) 350 800 350 800 Pie = 0 NOx Pie = 0 NOx Pie = 200 mbar Pie = 300 mbar 300 Pie = 400 mbar 700 300 Pie = 600 mbar 700 Pie = 600 mbar Pie = 900 mbar 250 600 250 600

200 500 200 500

150 400 150 400 Tem Tem 100 300 100 300

50 200 50 200

Equivalence ratio Equivalence ratio 0 100 0 100 0.0 0.2 0.4 0.6 0.8 1.0 1.2 0.0 0.2 0.4 0.6 0.8 1.0 1.2 Fig. 7 – NOx concentration and exhaust manifold Fig. 8 – NOx concentration and exhaust manifold temperature versus equivalence ratio φ for different temperature versus equivalence ratio φ for different relative engine inlet pressure at N = 1500 rpm relative engine inlet pressure at N = 2000 rpm

3.3.2 Hydrocarbon and Carbon Monoxyde: HC, CO HC and CO curves versus equivalence ratio are close (Figures 9 and 10). For the relative pressure Pie = 0, it is observed an increase of HC and CO at low equivalence ratio, especially for 2000 and 2500 rpm. This effect is probably due to the counter pressure at engine outlet. Pie = 0 means that boost pressure is equal to 0. In reality, this experiment corresponds to a test without supercharging, so there is a negative relative pressure at engine inlet and a positive pressure at engine outlet. Hence, we have a bad filling of the engine and exhaust gas recirculation.

7 HC (ppm) CO (ppm) HC (ppm) CO (ppm) 100 2000 100 2000

90 90 Pie = 0 Pie = 0 80 Pie = 200 mbar 1600 80 Pie = 250 mbar 1600 Pie = 400 mbar Pie = 500 mbar Pie = 600 mbar Pie = 750 mbar 70 70 Pie = 1000 mbar 60 1200 60 1200 HC HC 50 50 40 800 40 800 30 30 20 400 20 400 CO CO 10 10 Equivalence ratio Equivalence ratio 0 0 0 0 0.0 0.2 0.4 0.6 0.8 1.0 1.2 0.0 0.2 0.4 0.6 0.8 1.0 1.2 Fig. 9 – HC and CO concentration versus equivalence Fig. 10 – HC and CO concentration versus equivalence ratio φ for different relative engine inlet pressure ratio φ for different relative engine inlet pressure at N = 1500 rpm at N = 2500 rpm

3.3.3 Carbon Dioxyde: CO2 CO2 emissions are proportional to equivalence ratio and independent of boost pressure as indicated in Figures 11 and 12.

CO2 (%) CO2 (%) 16 16

14 Pie = 0 14 Pie = 200 mbar Pie = 0 Pie = 400 mbar Pie = 250 mbar 12 Pie = 600 mbar 12 Pie = 500 mbar Pie = 750 mbar Pie = 1000 mbar 10 10

8 8

6 6

4 4

2 2 Equivalence ratio Equivalence ratio 0 0 0.00.20.40.60.81.01.2 0.00.20.40.60.81.01.2 Fig. 11 – CO2 concentration versus equivalence ratio φ Fig. 12 – CO2 concentration versus equivalence ratio for different relative engine inlet pressure φ for different relative engine inlet pressure at N = 1500 rpm at N = 2500 rpm

3.4 Influence of boost pressure on specific gas emissions As noticed in chapter 3.1, for the same equivalence ratio brake specific fuel consumption improves when boost pressure increase and especially at low ratio. This effect has a direct influence on specific gas emission and we have approximately for all the gas considered: - same level of specific gas emission for equivalence ratio between 0.4 and 0.8 - lower level with increasing of boost pressure for equivalence ratio less than 0.4. These results are represented for 1500 rpm in Figures 13 to 16. 8 NOx (g/kw/h) CO (g/kw/h) 12 12

Pie = 0 Pie = 0 10 Pie = 200 mbar 10 Pie = 200 mbar Pie = 400 mbar Pie = 400 mbar Pie = 600 mbar Pie = 600 mbar 8 8

6 6

4 4

2 2

Equivalence ratio Equivalence ratio 0 0

0.0 0.2 0.4 0.6 0.8 1.0 1.2 0.0 0.2 0.4 0.6 0.8 1.0 1.2 Fig. 13 – NOx specific emission versus equivalence Fig. 14 – CO specific emission versus equivalence ratio φ for different relative engine inlet pressure ratio φ for different relative engine inlet pressure at N = 1500 rpm at N = 1500 rpm

HC (g/kw/h) CO2 (g/kw/h) 1.4 1600

Pie = 0 Pie = 0 1.2 Pie = 200 mbar Pie = 200 mbar Pie = 400 mbar 1400 Pie = 400 mbar Pie = 600 mbar Pie = 600 mbar 1.0 1200 0.8

1000 0.6

0.4 800

0.2 600 Equivalence ratio Equivalence ratio 0.0

0.0 0.2 0.4 0.6 0.8 1.0 1.2 0.0 0.2 0.4 0.6 0.8 1.0 1.2 Fig. 15 – HC specific emission versus equivalence Fig. 16 – CO2 specific emission versus equivalence ratio φ for different relative engine inlet pressure ratio φ for different relative engine inlet pressure at N = 1500 rpm at N = 1500 rpm

4 DISCUSSION Results show that boost pressure have no significant influence on emissions for the same equivalence ratio, except for NOx where there is a slight tendency of increase with boost pressure for engine speed 2000 and 2500 rpm. If specific emissions are considered, we find lower level of pollutants for low equivalence ratio, which is due to improvement of brake specific fuel consumption with the increase of boost pressure. This benefit is doubtful because in calculating Bsfc of the engine, it has not been taken into account the power required to produce the supercharging. In Figure 17 is represented the variation of Bsfc assuming an efficiency of compressor of 1.0 and 0.7 . So, it appears that this advantage can be cancelled by a drastic increase of Bsfc at low load for high boost pressure. It is logical because the power needed to produce the compressed air become equivalent to the power produced by the engine (about 2.5 kW are necessary to get an inlet engine pressure of 600 mbar at 1500 rpm). 9 Bsfc (g/kw.h) 500

450

Pie = 600 mbar 400

350 η

is = 0 . η 7

i s = 300 1 .0

250

Torque (N.m) 200 0 40 80 120 160 200 240 280 Fig. 17 – Bsfc versus torque for compressor efficiency 0.7 and 1.0 at N = 1500 rpm

CONCLUSION Increases of boost pressure at low engine speeds don't lead directly to significant improvement of pollutant emissions or brake specific fuel consumption. However, opportunities can be found in supercharging: - supercharged engine has a higher torque, so regarding pollution vehicle can be more easily use in a convenient functioning area. - instead of using compressed air for increasing engine power, it can be forwarded to use this air for treatment of exhaust gas as for example post combustion in order to improve the catalyst or burn of particulate. This turbocharging concept oriented towards decreasing pollution emissions is already used in Diesel engines and it’s now the object of interest for new direct injection engines.

REFERENCES 1. Anada S, Kawakami T, Shibata N. Development of swirl jet turbocharger for Diesel engine vehicles. SAE paper 970341. 2. Toussaint M, Descombes G, Pluviose M. Research into variable geometry turbochargers without wastegate. ImechE 3rd European Turbomachinery Conference, London U.K. March 1999. Paper C557/11ASME Fall Technical Conference Ann Arbor 1999 Paper n° 99-ICE-228 ICE-Vol. 33-2 3. Descombes G, Jullien J. A new computer modelling for variable nozzle on an advanced supercharging engine. 45th ASME International Gas Turbine Conference, Munich Germany May 2000. 4. Podevin P, Descombes G, Marez P, Dubois F. A Study of turbocharged Diesel engine during sudden acceleration. Set up and exploitation of a test rig. ASME Fall Technical Conference, Ann Arbor USA October 1999. Paper n° 99-ICE-228 ICE-Vol. 33-2 5. Magnet J.L, Thermoenergetic engines, cours de Moteurs du CNAM1997-1998