The Advanced Engineering Centre (AEC), led by Professor Morgan Heikal, has an international reputation for producing innovative, future-facing research.
2003-01-0750
Evaluation of HCCI for Future Gasoline Powertrains R J Osborne, G Li, S M Sapsford, J Stokes, and T H Lake Ricardo
M R Heikal University of Brighton Copyright © 2003 Society of Automotive Engineers, Inc.
ABSTRACT This paper describes a two-year programme of research conducted by the authors investigating HCCI in direct injection gasoline engines. Poppet-valved two-stroke cycle operation has been investigated experimentally, using conventional gasoline compression ratios and fuel, and ambient temperature intake air. Extensive combustion and emissions data was gathered from the experimental engine. Computational Fluid Dynamics (CFD) has been used to model HCCI combustion, and the CFD tool validated using experimental data. Based on experience with the two-stroke engine and modelling techniques, a four-stroke engine has been designed and tested. Using this range of tools, practical options for gasoline HCCI engines are evaluated, and a scenario for the market introduction of HCCI is presented.
INTRODUCTION Homogeneous charge compression ignition (HCCI) is best regarded as a combustion mode of internal combustion engines distinct from the conventional spark ignition (SI) and compression ignition (CI) operating modes. In an HCCI engine, a homogeneous mixture of air, fuel and residual gases is compressed until autoignition occurs at sites distributed throughout the combustion chamber. The charge is then consumed by controlled autoignition reactions. In the ideal case, there is no occurrence of either the propagating flame front that characterises SI combustion, or the diffusion burning that characterises CI combustion. HCCI combustion in IC engines potentially offers benefits compared with both SI and CI combustion. Indeed, it may be seen to combine the advantages of the two conventional combustion modes. Through the use of very dilute mixtures an HCCI engine could operate unthrottled at part load – reducing pumping losses – as the diesel engine does. The overall burn rates of HCCI combustion
are typically fast [1], and if correctly phased in the cycle could approximate the ideal Otto cycle. The combustion processes that characterise HCCI are found to result in significantly lower temperatures than those inside the reaction zone of an SI engine. As a result the NOx emissions from HCCI engines are dramatically reduced. The unburned hydrocarbon emissions from HCCI engines are a subject of sustained debate, but many workers have reported HC emissions to be higher than for comparable CI and SI engines [2,3]. The authors have adopted the term Homogeneous Charge Compression Ignition (HCCI) throughout this paper to identify the combustion processes investigated. The clear drawback of this designation concerns the degree of homogeneity; in many cases the charge may be significantly heterogeneous. As such, the authors have sympathy with research groups using alternative designations, most notably CAI (Controlled Auto-Ignition) [4,5]. However, it is felt that HCCI is appropriate when there is no deliberate stratification of fuel, and the term has been adopted by a number of national and international institutions.
PROGRAMME APPROACH The objective of the research investigation has been to demonstrate HCCI in two-stroke and four-stroke DI gasoline engines, and through doing so to develop a greater understanding of the processes that constitute HCCI. A wide range of tools have been employed to achieve these objectives, including single-cylinder engine testing, optical engine testing, one-dimensional gas dynamics simulation, and three-dimensional computational fluid dynamics (Figure 1). From the previous experience of the authors and the literature available, it was clear that HCCI combustion could be more readily stimulated in a two-stroke cycle engine. Ricardo have significant previous experience in poppet-valved two-stroke engines, developing the
Flagship engine concept that has been used for the twostroke HCCI engine in this investigation. Two-stroke engines Conventional engines
Optical engines
Modelling tools WAVE
1-D
Four-stroke engines
VECTIS
3-D
Conventional engines
Optical engines
Combustion system design Figure 1. Programme approach
Having successfully recorded HCCI combustion in a twostroke engine, the in-cylinder conditions were further explored and characterised using a combination of 1-D gas dynamics and 3-D CFD computations, along with optical-access studies. When confidence in the incylinder conditions needed for HCCI combustion was established, the same tools could be employed to design a four-stroke HCCI combustion system. A four-stroke HCCI single-cylinder engine was subsequently constructed, tested and developed.
TWO-STROKE ENGINE TESTING
The intake ports are inclined vertically as shown in Figure 2, and boost air is applied in order to achieve a loop scavenging action. The great advantage of using the Flagship engine in this investigation is that the two-stroke and four-stroke engine combustion chambers are nearly identical. Further details of the two-stroke engine are presented in Table 1, along with a photograph of the testbed installation (Figure 3). Standard pump gasoline was used throughout this study, and the engine was fitted with production-type DI gasoline FIE. The testbed intake and exhaust systems were constructed from 50mm diameter pipe, with a compressed air source available for boosting. A two-stroke engine tuned exhaust system was not used. The engine was equipped with a throttle, but it was maintained in the wide-open position for all twostroke engine tests. Table 1. Two-stroke engine specification Bore
74.0 mm
Stroke
75.5 mm
Swept Volume
324.7 cc
Compression ratio (geometric)
9
Piston
Flat
Fuel
BP 95 RON ULG
Fuel pressure
100 bar
The Flagship engine concept was developed at Ricardo between 1989 and 1992 as a two-stroke engine with a high degree of commonality with modern four-valve fourstroke gasoline engines [6,7]. The Flagship engine thus employs four poppet valves, a central spark plug and an in-cylinder fuel injector mounted at the edge of the combustion chamber between the intake valves.
Figure 3. Single-cylinder engine testbed installation
Figure 4 illustrates the valve timing employed by the Flagship engine. In this investigation valve events were selected in order to yield the high residual rates ideal for HCCI combustion.
Figure 2. Flagship engine layout
Cumulative normalised heat release [%]
TDC
EVO IVC
CA EVC IVO
SOI BDC
120 100
2500 rev/min 2.9 bar IMEP 19:1 AFR
80 60 40 20 0 -20 -60
-30
0
30
60
Crank angle [deg]
Figure 4. Typical valve timing diagram for Flagship two stroke HCCI engine Rate of heat release [%/°CA]
A comprehensive test programme was carried out using the two-stroke single-cylinder engine, investigating the influence of all engine parameters on HCCI combustion. Figures 5 and 6 – PV diagram and heat release plots – illustrate a typical HCCI operating point with the spark plug inactive. The engine speed is 2500 rev/min and overall air-fuel ratio of 19:1 was measured in the exhaust stream. Trapped and overall air-fuel ratios can of course differ significantly in the two-stroke cycle. The rapid combustion characteristic of HCCI is evident, with peak heat release occurring close to TDC, and a 10 - 90% mass fraction burned (MFB) duration of only 8.2 °CA.
14 2500 rev/min 2.9 bar IMEP 19:1 AFR
12 10 8 6 4 2 0 -2 -60
-30
0
30
60
Crank angle [deg]
35 2500 rev/min 2.9 bar IMEP 19:1 AFR
Cylinder pressure [bar]
30
Figure 6. Heat release diagrams for a typical two-stroke HCCI operating condition
HCCI combustion was recorded with a wide range of mixture strength. For an engine speed of 2750 rev/min and an inlet manifold pressure fixed at 250 mbarg, Figures 7 and 8 illustrate the influence of air-fuel ratio on heat release, fuel consumption and emissions. The limits of HCCI combustion – at 13.5:1 and 26.5:1 air-fuel ratio – were defined by combustion stability, with an operating limit of 5% for coefficient of variation (COV) of IMEP applied.
25
20
15
It is observed that the most rapid heat release occurs with an overall air-fuel ratio of 20:1, and that the angle of 50% MFB occurs significantly before TDC in this case. Subsequent engine test results are presented for an overall AFR of 22:1, chosen for minimum fuel consumption and HC emissions, although NOx emissions are close to maximum for this condition.
10
5
0
0
1
2
3
4
5
6
7
8
9
10
Normalised cylinder volume Figure 5. PV diagram for a typical two-stroke HCCI operating condition
2.8
2750 rev/min Inlet manifold 250 mbar g
2.6
IMEP [bar]
2.4 2.2 2.0 1.8 1.6 1.4 15 10 5 0 -5 -10 -15 -20
5.0
22:1 AFR 4.5
15 12 9
3.0
3
1.5 14
16
18
20
22
24
26
Valvetrain limit
2.5 2.0
12
Misfire limit
1.0
28
Air-Fuel Ratio (Spindt)
0.5 1800
2000
Figure 7. Effect of AFR on heat release in two-stroke engine
2750 rev/min Inlet manifold 250 mbar g
2.4 2.2 2.0 1.8
500 450 400 350 300
1.2
ISFC [g/kWh]
1.6
2400
2600
2800
3000
3200
3400
Engine speed [rev/min]
Figure 9 illustrates the operating regions recorded with a fixed air-fuel ratio, and those employing all possible airfuel ratios. In this investigation two-stroke operation was limited to a maximum engine speed of 3250 rev/min by valvetrain considerations. There is no reason to believe that HCCI operation would not be possible at significantly higher engine speeds. Figure 10 records the inlet manifold pressures required to achieve the HCCI loads, for the maximum operating speed of 3250 rev/min. 4.5
250
4.0
1.0
3.5 IMEP [bar]
0.8 0.6 0.4 0.2
50 40 30 20 10 12
14
16
18
20
22
24
26
28
2.5
3250 rev/min 22:1 AFR
1.5 1.0 100
200
300
400
500
Inlet manifold pressure [mbarg]
0
Air-Fuel Ratio (Spindt)
Figure 8. Effect of AFR on emissions and fuel economy in two-stroke engine
3.0
2.0
ISHC [g/kWh]
IMEP [bar]
2200
Figure 9. Operating regions for two-stroke HCCI operation
2.6
ISNOx [g/kWh]
Scavenge limit
3.5
6
0
All air-fuel ratios
4.0
IMEP [bar]
10-90% MFB [°CA]
18
50% MFB [°CA ATDC]
1.2
For a single fixed valve timing, a significant operating region of HCCI was recorded. There are two clear boundaries for HCCI operation with the two-stroke cycle. At higher loads scavenging efficiency increases to a point at which there is insufficient residual retained to allow autoignition. As engine speed increases, the scavenging efficiency for a fixed load decreases, so HCCI is possible at higher loads with greater engine speed. At low loads large proportions of residual are trapped, but the exhaust temperature is decreasing and a limit is reached at which autoignition is not possible due to low overall charge temperature.
Figure 10. Effect of boost pressure on load with the Flagship engine
The combustion and emissions performance of the engine in the HCCI operating mode are shown in Figures 11 to 17, with fixed overall air-fuel ratio. Indicated values
are calculated on a 360 °CA basis. By this definition the two-stroke engine will produce twice the torque for a particular IMEP compared with a four-stroke engine. 4.5
4.5
ISCO [g/kWh]
4.0
4.0
3.5
3.5
IMEP [bar]
IMEP [bar]
ISFC [g/kWh]
260
3.0
265 270 280
2.5 300
275
290
12 8
1.5
1.5
2200
2400
2600
2800
3000
3200
1.0 1800
3400
4
2.5
2.0
2000
16
3.0
2.0
1.0 1800
8
4
1
8
2000
2200
2400
2600
2800
3000
3200
3400
Engine Speed [rev/min]
Engine Speed [rev/min]
Figure 14. Two-stroke engine ISCO emissions [g/kWh] for exhaust AFR fixed at 22:1 overall
Figure 11. Two-stroke engine ISFC [g/kWh] for exhaust AFR fixed at 22:1 overall
4.5
4.5
Exhaust temperature [°C]
ISNOx [g/kWh] 4.0
4.0
3.5
3.5
IMEP [bar]
IMEP [bar]
600
3.0
3.0
2.4 2.0 1.6
2.5
570 540
3.0
510 480
2.5
450
1.2
2.0
420
2.0
0.8
390
1.0 1800
360
0.4
1.5
1.5
0.2
2000
2200
2400
2600
2800
3000
3200
1.0 1800
3400
2000
2200
2400
2600
2800
3000
Engine Speed [rev/min]
3400
Figure 15. Two-stroke engine exhaust temperature [°C] for exhaust AFR fixed at 22:1 overall
Figure 12. Two-stroke engine ISNOx emissions [g/kWh] for exhaust AFR fixed at 22:1 overall
4.5
4.5
Coefficient of Variation of IMEP [%]
ISHC [g/kWh] 4.0
4.0
3.5
5
3.5
6.5
IMEP [bar]
6
IMEP [bar]
3200
Engine Speed [rev/min]
3.0 24
19 15 12
2.5
9
5.0 4.0
3.0
3.0
2.5 2.5
2.0
2.0
5
6
1.5
12
2000
2200
3.0
9
1.5
1.0 1800
5.0 4.0
2400
2600
2800
3000
3200
Engine Speed [rev/min]
Figure 13. Two-stroke engine ISHC emissions [g/kWh] for exhaust AFR fixed at 22:1 overall
3400
1.0 1800
4.0
2000
2200
2400
2600
2800
3000
3200
Engine Speed [rev/min]
Figure 16. Two-stroke engine coefficient of variation of IMEP [%] for exhaust AFR fixed at 22:1 overall
3400
where
4.5
Angle of 50% Mass Fraction Burned [°CA ATDC(F)]
T t ig = AP − n exp A T
4.0
IMEP [bar]
3.5
Combustion takes place when
3.0 9 9 3 3
0
0
-3
-3
2.0 3
1.5
3
2000
2200
2400
2600
2800
3000
3200
Engine Speed [rev/min]
Figure 17. Two-stroke engine angle of 50% mass fraction burned [°CA ATDC(F)] for exhaust AFR fixed at 22:1 overall
CFD SIMULATION The main purpose of performing CFD simulation in this study was to reveal the critical in-cylinder conditions required to stimulate HCCI. Conditions such as incylinder charge temperature and residual fraction are of course difficult to determine experimentally. In addition, to support one-dimensional engine cycle simulation the detailed two-stroke scavenging characteristics were produced through CFD. CFD CODE AND MODELS The Ricardo engine CFD code VECTIS was used in this study. VECTIS is a finite volume based code solving for the three-dimensional flow equations governing the conservation of mass, momentum and energy. Features of automatic mesh generation and time dependent mesh motion are used to cope with engine flow with moving piston and valves. The fuel spray is calculated using the discrete droplet method (DDM). The spray is represented by an ensemble of droplet parcels which are introduced at the injector with initial conditions determined from a specified injection rate, spray angle and droplet size distribution. Sub-models of secondary break-up, droplet-turbulence interaction, droplet-droplet interaction and impingement are incorporated, accounting for the subsequent behavior. The Reitz-Diwakar break-up model [8] was used in this study for secondary break-up.
3400
The combustion model used in this work was the Ricardo Two-Zone Flamelet (RTZF) model [10]. The RTZF model is a simplified coherent flamelet combustion model based on a two-zone representation. Each computational cell is notionally divided into burned and unburned zones, and the unburned zone is further divided into segregated and mixed regions. In non-premixed combustion, fuel/air mixing is continuously calculated based on the eddy break-up concept. The mixed reactants are transferred into the mixed region where they accumulate until combustion takes place. The combustion products generated during combustion are transferred into the burned zone. The RTZF model calculates the combustion rate based on the joint effect of chemical kinetics controlled burning and normal turbulence controlled burning [11]. CASES STUDIED Four HCCI cases and one SI case were selected from the two-stroke engine test programme for CFD simulation. The case conditions are summarized in Table 2 and the operating conditions illustrated in Figure 18. Each of the cases, including the SI condition, uses the same valve timing. Table 2. Two-stroke simulation cases
Case number
A
B
C
D
S
Engine speed [rev/min]
2250
2753
3255
3258
1500
IMEP [bar]
1.967
2.61
3.834 1.566
3.91
Overall AFR
19.39 18.67 19.17 19.15 20.82
Fuel injection mass [mg]
5.74
7.14
9.85
4.47
10.51
180
180
180
180
200
200
214
o
Nominal SOI [ CA ATDC] o
Auto-ignition is predicted by the Livengood-Wu integral [9]. The ignition is controlled by a probability function
Nominal EOI [ CA 193.5 ATDC] Operating mode
Pig = ∫
dt t ig
Pig reaches unity.
6 6
2.5
1.0 1800
t ig is the ignition delay time defined as
201.5 217.5
HCCI HCCI HCCI HCCI
SI
4.5
IMEP [bar]
4.0
S SI Operation
The uncertainties regarding the trapped conditions in the two-stroke engine simulation were resolved by a multidomain, multi-cycle simulation process. Simulation starts with a estimated initial condition. The complete scavenging, fuel injection and combustion processes were then solved repeatedly from cycle-to-cycle until fully developed, and the initial uncertainties were progressively removed. The multi-domain algorithm allows disconnected domains to be solved simultaneously. The inlet and/or exhaust ports behind the closed valves were kept in the simulation and all of the information retained.
C
3.5
HCCI Operation
3.0
B
2.5
A
2.0
D
1.5 1.0 1000
1500
2000
2500
3000
3500
SIMULATION RESULTS
Engine Speed [rev/min]
SIMULATION PROCEDURE Prior to the CFD simulation, Cartesian based computational meshes were generated in the preprocessing phase of the CFD code. The input was the complete surface model including the combustion chamber and intake and exhaust ports in STL format (Figure 19). The movement of the piston and the intake and exhaust valves were resolved using the boundary motion feature of the CFD code. During the solution, as the boundaries move, the internal mesh structure deforms automatically to minimize the distortion of cells. When the general cell distortion reaches a certain level the solution is re-zoned onto a new set of Cartesian mesh.
Figure 20 shows the comparison of the measured and simulated cylinder pressures. With the ignition model scaling factor tuned for the first case and kept constant in all cases, the ignition point in most HCCI cases was well captured. The only exception is in case C where the predicted ignition point was about 8 °CA earlier than the measured. The peak pressure and the pressure gradient over the combustion period produced by CFD simulation match closely with the measurement. For the spark ignition condition (case S), the predicted ignition and pressure gradient over the combustion phase agreed extremely well with the measurement, but the peak cylinder pressure was underpredicted by about 2 bar. 25
Experiment Cylinder pressure [bar]
Figure 18. Two-stroke CFD case operating conditions
VECTIS
20
15
10
5
0 -180
-90
0
90
180
Crank angle [deg]
(a) Case A
Cylinder pressure [bar]
35 Experiment
30
VECTIS
25 20 15 10 5 0 -180
-90
0
90
180
Crank angle [deg]
(b) Case B Figure 19. Engine geometry for CFD calculations
Figure 20. Comparison of the measured and predicted incylinder pressures (continued on next page)
(2) The phenomenon of charge reversal is captured by the multi-domain, multi-cycle CFD simulation. The cylinder contents, including fresh fuel, air and residuals, are pushed back into the inlet ports while the inlet valves are still open and the piston starts to move up (seen at o 220 CA). This part of the mixture will be fed back into the cylinder at the start of the next intake stroke. Part of this will be short-circuited into the exhaust ports (seen at 160 o CA).
Cylinder pressure [bar]
50
Experiment VECTIS
40
30
20
10
0 -180
-90
0
90
180
Crank angle [deg]
(c) Case C
Cylinder pressure [bar]
30
Experiment VECTIS
25
20 15 10
5 0 -180
-90
0
90
180
Crank angle [deg]
(d) Case D
(3) Charge inhomogeneity, in terms of temperature and mixture composition, remains all the way through the o compression stroke (seen at 220, 280, 340 CA). It is interesting to see that the auto-ignition starts around the cylinder centre closer to the intake side (seen at 360 o CA). Figures 22 and 23 compare the in-cylinder charge o temperature at 20 CA BTDC and the trapped residuals from the CFD simulations for HCCI and SI cases. Under the operating condition of case S, when HCCI combustion could not be achieved, the charge temperature is significantly lower than that in any of the HCCI cases. The residual level in case C is not much higher than in case S. However, the higher load condition results in higher overall in-cylinder temperature which enables HCCI.
30
0.7
VECTIS 0.6
20
Residual fraction
Cylinder pressure [bar]
Experiment 25
15 10
5 0 -180
0.5 0.4 0.3 0.2 0.1
-90
0
90
180
Crank angle [deg]
0
A
(e) Case S Figure 20. Comparison of the measured and predicted incylinder pressures (continued)
B
C
D
S
CFD cases
Figure 22. Two-stroke engine residual fractions from CFD for SI and HCCI cases
Figure 21 demonstrates the in-cylinder processes reproduced by CFD simulation using a series of distribution plots illustrating velocity, fuel mass fraction, temperature and combustion products/residuals across the cylinder centre plane and valve centre plane. A number of features are revealed: (1) The two-stroke scavenging process is suppressed with the chosen valve timing strategy. It is clear that a large amount of residuals is retained inside the cylinder at the time when the exhaust valves are closed (seen at o 220 CA). The trapped residuals are accounted for in modifying the in-cylinder conditions in order to achieve HCCI.
Mean charge temperature at 20 °CA BTDC [K]
1300 1200 1100 1000 900 800 700 A
B
C
D
S
CFD cases
Figure 23. Two-stroke engine charge temperatures from CFD for SI and HCCI cases
Velocity
Fuel mass fraction
Temperature
Products mass fract.
Velocity
Crank angle = 140
Crank angle = 160
Crank angle = 200
Crank angle = 220
Fuel mass fraction
Figure 21. Two-stroke engine case A (continued on next page)
Temperature
Products mass fract.
Velocity
Fuel mass fraction
Temperature
Products mass fract.
Velocity
Crank angle = 280
Crank angle = 340
Crank angle = 360
Crank angle = 80
Fuel mass frac.
Figure 21. Two-stroke engine case A (continued)
Temperature
Products mass fract.
FOUR-STROKE ENGINE TESTING
40
Using results from CFD and one-dimensional simulation work, a four-stroke HCCI single-cylinder engine was designed. The differences between the two-stroke and four-stroke engines concern the piston and the valvetrain only, as shown in Table 3.
35
Pressure [bar]
30
Table 3. Four-stroke engine specification 74.0 mm
Stroke
75.5 mm
Swept Volume
324.7 cc
Compression ratio
11.7
Piston
Offset bowl
Fuel
95 RON ULG
Fuel pressure
100 bar
20 15
5
0
A Ricardo DI gasoline piston was used, featuring an offset bowl. This piston was used to enable a compression ratio suitable for modern DI gasoline engines, and to allow stratification of fuel if needed. The library of camshafts shown in Figure 24 were designed and procured. The profiles are representative of those that could be achieved with a fully flexible mechanical VVA system [12]. 120
Normalised valve lift [%]
25
10
100 80 60 40
0
2
6
8
10
12
Figure 25. PV diagram for a typical four-stroke HCCI operating condition 120 100
3500 rev/min 2.6 bar IMEP Stoichiometric AFR
80 60 40 20 0 -20 -60
-30
20 0 150
4
Normalised cylinder volume
Cumulative normalised heat release [%]
Bore
3500 rev/min 2.6 bar IMEP Stoichiometric AFR
0
30
60
Crank angle [deg] 200
250
300
16
Cam angle [deg]
As for the two-stroke engine, a test programme was undertaken to explore HCCI combustion with the fourstroke cycle. Figures 25 and 26 illustrate four-stroke HCCI, for an engine speed of 3500 rev/min and load of 2.6 bar IMEP. Ambient temperature intake air, standard gasoline fuel and a stoichiometric mixture were used. The in-cylinder temperatures needed for autoignition are delivered by trapped residual using the well-established negative overlap approach [13-16]. The re-compression event occurring at non-firing TDC that results from the early EVC and late IVO strategy is clearly seen in the PV diagram. As with the two-stroke cycle, heat release is rapid for HCCI combustion.
Rate of heat release [%/°CA]
Figure 24. Library of camshafts designed for four-stroke HCCI engine
3500 rev/min 2.6 bar IMEP Stoichiometric AFR
14 12 10 8 6 4 2 0 -2 -60
-30
0
30
Crank angle [deg]
Figure 26. Heat release diagrams for a typical four-stroke HCCI operating condition
60
Figure 27 shows the operating ranges tested using a stoichiometric mixture. Full HCCI was recorded above 2500 rev/min, with a partial HCCI condition at lower speeds. Test results were gathered between 1000 and 2500 rev/min with the spark plug active, and as is demonstrated by the later figures, there is no discontinuity in behaviour between full HCCI operation and spark-supported HCCI. 6.0 5.5
It is not possible to develop a complete hypothesis for the spark-supported condition using the engine test data alone, but it is hoped that current optical engine studies and CFD simulation will provide further information on this mode.
IMEP [bar]
5.0 4.5 4.0
Spark-supported HCCI
3.5
The figure shows the residual rates predicted by the onedimensional gas dynamics code WAVE for these conditions. The values of around 55% are significantly higher than the residual tolerance of this combustion system for SI operation. Heat release rates are fairly fast (10-90% MFB duration in the range 20 – 22 °CA), suggesting the occurrence of a distributed process. This may be contrasted with conventional SI operation using standard camshafts in this engine. For the same speed and load condition a burn duration of 35 °CA was recorded for a residual rate of 27%.
Full HCCI
3.0
6.0 5.5
2.5
1000
1500
2000
2500
3000
3500
4000
Engine speed [rev/min] Figure 27. Operating regimes for HCCI operation with stoichiometric mixture
Figure 28 illustrates the influence of spark timing on heat release in the spark-supported HCCI mode. Whilst combustion phasing broadly responds to the timing of the spark event, it is clear that there are significant differences between this mode and conventional SI operation.
3.00 2.95 12 11 10 9
10-90% MFB [°CA]
8 23.0
7
22.5 22.0
Angle of 50% MFB [°CA ATDC]
3.05
2.90
21.5 21.0 20.5
65 60 55 50 45
310
315
320
325
330
335
4.5 4.0
Boosted HCCI
3.5 3.0 2.5 2.0 500
1000
1500
2000
2500
3000
3500
4000
Engine speed [rev/min] Figure 29. Boosted HCCI operating region
2000 rev/min Stoichiometric AFR
3.10
40 340
Residual fraction [%]
IMEP [bar]
3.15
IMEP [bar]
5.0
2.0 500
Ignition timing [°CA ATDC (NF)]
Figure 28. Effect of ignition timing in spark-supported HCCI region
The testbed was equipped with a compressed air supply, and by raising the intake manifold pressure it was possible to achieve full HCCI across the speed range, as shown in Figure 29. Turning to the operating limits for HCCI combustion, the lower load limit is identical to that with two-stroke cycle operation: exhaust temperature decreases to the point at which misfire occurs. For the naturally aspirated full HCCI condition the upper load limit is also analogous to that with the two-stroke: it is not possible to trap sufficient residual to achieve autoignition. For boosted HCCI load is instead limited by maximum rate of pressure rise. Figure 29 shows data up to a maximum rate of pressure rise of 8 bar/°CA, which is very high by SI engine standards. For the spark-supported HCCI mode there is no clear upper operating limit, but rather a gradual transition to conventional SI operation. Figures 30 to 35 illustrate the characteristics of the fourstroke engine with ambient pressure intake air and a stoichiometric mixture. Indicated results are shown as net values, calculated on a 720 °CA basis.
6.0
6.0
ISCO [g/kWh]
ISFC [g/kWh] 5.5
5.5
5.0
5.0
285
IMEP [bar]
IMEP [bar]
35
4.5 4.0 3.5
285
320
270
260
270
4.5 35
4.0 3.5
3.0
3.0
2.5
2.5
2.0 500
1000
1500
2000
2500
3000
3500
2.0 500
4000
35
35
60
20
12 35
60
1000
20
1500
Engine speed [rev/min]
2000
2500
3000
3500
4000
Engine speed [rev/min]
Figure 30. Four-stroke engine ISFC [g/kWh]
Figure 33. Four-stroke engine ISCO emissions [g/kWh]
6.0
6.0
ISNOx [g/kWh]
Exhaust temperature [°C]
5.5
5.5
5.0
5.0
3.5
4.5
IMEP [bar]
IMEP [bar]
2.5 1.4
4.0
0.8 0.5
3.5
4.5 4.0 400
3.5 310
3.0
430
460
490
370
340
3.0
0.2 0.1
2.5 2.0 500
2.5
1000
1500
2000
2500
3000
3500
2.0 500
4000
1000
1500
Engine speed [rev/min]
2000
3000
3500
Figure 31. Four-stroke engine ISNOx emissions [g/kWh]
Figure 34. Four-stroke engine exhaust temperature [°C]
6.0
6.0
ISHC [g/kWh] 5.5
5.0
5.0
4.5 6
4.0 8
3.5
10
3.0
4000
Coefficient of Variation of IMEP [%]
5.5
IMEP [bar]
IMEP [bar]
2500
Engine speed [rev/min]
4.5 4.0
3.0
13
2
1
3
3.5 5
3
2
18
2.5 2.0 500
2.5
1000
1500
2000
2500
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3500
Engine speed [rev/min]
Figure 32. Four-stroke engine ISHC emissions [g/kWh]
4000
2.0 500
1000
1500
2000
2500
3000
3500
Engine speed [rev/min]
Figure 35. Four-stroke engine coefficient of variation of IMEP [%]
4000
IN-CYLINDER CONDITIONS Residual-promoted autoignition remains the most practical realisation of HCCI combustion. Retained exhaust products are used to provide the thermal energy necessary for autoignition, with stoichiometric or nearstoichiometric trapped air-fuel ratio. Through the combination of experimental and modelling studies, an increasingly detailed understanding of the incylinder conditions required for HCCI combustion has been developed. In this study the circle of experiment and model validation has been employed to characterise the required charge conditions, and confirm the transference of principles between two-stroke and fourstroke cycle operation. For both two-stroke and four-stroke operation, residual fractions in the range 40 – 70% were required to stimulate HCCI, as shown in section 4.4. These charge conditions produced mean gas temperatures in the range 950 - 1150 K at 20 °CA prior to TDC.
conclusions may be somewhat different. As shown in Figure 37, the HCCI mode does indeed show an increase in HC emissions by comparison with homogeneous DI operation with a stoichiometric mixture. However, the HC emissions for HCCI combustion are lower than for stratified charge DI operation, and for homogeneous operation with a lean mixture of λ = 1.4. Zero smoke emissions were measured in all cases in both the two-stroke and four-stroke engines. 350
2000 rev/min 2.7 bar IMEP 284
300
ISHC emissions [%]
ANALYSIS AND DISCUSSION
250 200
177 138
150 100 100 50 0
Stoichiometric Stoichiometric DI HCCI
Stratified DI
Lean DI lambda 1.4
Figure 37. Comparison of HC emissions for SI and HCCI combustion at 2000 rev/min, 2.7 bar IMEP
EXHAUST EMISSIONS The single largest attraction of HCCI combustion for application in future automotive powertrains is very low emissions of nitrogen oxides. In this study ISNOx emissions in the range 0.1 to 3 g/kWh were recorded, offering a reduction of 50 – 99% compared with a representative SI engine. Figure 36 illustrates a key point comparison for HCCI operation and different operating modes for the baseline DI gasoline engine. Data for the SI modes was gathered using fixed camshafts and without EGR dilution. 160
2000 rev/min 2.7 bar IMEP
146
FUEL CONSUMPTION Fuel economy is also considered an attractive feature of HCCI combustion. Figure 38 shows the net ISFC results for the four operating modes considered. HCCI combustion with a stoichiometric mixture demonstrates an 8% improvement in ISFC compared with the baseline stoichiometric SI case. This benefit is comparable to that recorded with homogeneous lean burn operation, but less than that for stratified charge operation. Fuel consumption in the HCCI mode can be further reduced by using a lean mixture.
120
110
100
2000 rev/min 2.7 bar IMEP
100 100 100
80 60 44 40 20 1
Net ISFC [%]
ISNOx emissions [%]
140
92
91
90 83 80
0
Stoichiometric Stoichiometric DI HCCI
Stratified DI
Lean DI lambda 1.4
Figure 36. Comparison of NOx emissions for SI and HCCI combustion at 2000 rev/min, 2.7 bar IMEP
Most recent research in the field has reported that unburned hydrocarbon emissions from HCCI engines are higher than those from modern four-stroke gasoline and diesel engines. As always however, the baseline used for comparison is critical. When assessing the HCCI mode compared with SI results for the base direct injection gasoline combustion system operating modes, the
70
60
Stoichiometric Stoichiometric DI HCCI
Stratified DI
Lean DI lambda 1.4
Figure 38. Comparison of net ISFC for SI and HCCI combustion at 2000 rev/min, 2.7 bar IMEP
Consideration of the factors that contribute to the net ISFC results is instructive. Figure 39 shows the PMEP comparison for the four modes. The negative overlap HCCI strategy realizes a significant pumping work
reduction compared with both stoichiometric and homogeneous lean SI operation. (The stratified charge condition of course has near-zero pumping losses.) This result means that the gross ISFC for the HCCI case is higher than for lean burn SI and close the stoichiometric SI case, a result that was not anticipated. 0.70
2000 rev/min 2.7 bar IMEP 0.60
0.56 0.49
PMEP [bar]
0.50 0.40
0.36
0.30
Figure 40 presents a possible scenario for the staged introduction of HCCI into gasoline powertrains, with first and second generation HCCI engines and the future vision of a two-stroke /four-stroke switching engine. A fundamental issue is that operating ranges for HCCI remain modest, and HCCI engines will therefore be dual mode in nature, using SI combustion for starting, higher loads and very low loads. This imposes considerable constraints on what may be changed in the engine specification to assist HCCI, in particular dictating the use of standard fuel and conventional compression ratio. It is also unlikely that intake air heating could be applied in an HCCI-SI engine.
0.20 0.08
0.10 0.00
Stoichiometric Stoichiometric DI HCCI
Stratified DI
Lean DI lambda 1.4
Figure 39. Comparison of PMEP for SI and HCCI combustion at 2000rev/min, 2.7 bar IMEP
A further issue in the development of practical HCCI powertrains is overlap between HCCI and SI operation. If HCCI combustion falters for any reason, it is critical that combustion is quickly restored by the spark plug. The authors have sought forms of HCCI in which there is significant overlap with SI operation. The subsequent sections address particular aspects of future HCCI engine design and control.
APPLICATION OF HCCI IN GASOLINE POWERTRAINS
VALVETRAIN REQUIREMENTS With a wealth of experimental and modelling data for HCCI combustion now available, it is appropriate to address the role of HCCI in future passenger car engines. It is the view of the authors that an ‘HCCI revolution’ – with wholesale changes made to engine specifications in order to accommodate HCCI – is unlikely to take place. More plausibly, as the enabling technologies for HCCI are introduced and combined, the particular benefits of HCCI can gradually be employed to meet legislative and market requirements.
First-Generation HCCI
Valvetrain selection is critical for HCCI engines, as the valve events help to deliver the charge conditions needed for autoignition. This investigation presents HCCI combustion that could be practically realised using a fully flexible mechanical VVA system – of the type now in series production applied to both the intake and exhaust valves. This type of valvetrain is realistic for a first-generation implementation of HCCI (Figure 40). Second-Generation HCCI
2S-4S Switching Engine
Model Year Combustion system
Any
Any
Two-stroke capable
Standard
Standard
Standard
Conventional SI
Conventional SI
Conventional SI
DI likely
DI likely
DI
Valvetrain
Mechanical VVA
Camless
Camless
Aspiration
NA
NA
Electric assist turbo
Conventional
CPEMS
CPEMS
TWC
TWC
TWC and LNT
Fuel Compression ratio FIE
EMS Aftertreatment
Increasing flexibility
Figure 40. Introduction of HCCI powertrains
MIXTURE PREPARATION This investigation has applied in-cylinder injection of fuel in all cases. Under laboratory conditions there is no reason to believe that direct injection is required for HCCI combustion. Indeed, the greater mixture homogeneity offered by port fuel injection may offer some benefits in terms of exhaust emissions and combustion stability. However, there are several reasons why direct injection may prove essential in the practical application of gasoline HCCI. Foremost among these is the requirement to switch frequently between SI and HCCI operation. The flexibility of mixture preparation offered by DI will be a significant advantage in smooth mode switching. AFTERTREATMENT REQUIREMENTS As discussed in previous sections, the HCCI engine data gathered indicates very low NOx emissions, and HC and CO emission levels no greater than those from other DI gasoline engine modes. As a result of the limited operating ranges achieved with HCCI, both test cycle and real world driving would necessitate significant periods spent using SI operation. As such, the exhaust aftertreatment specification for an HCCI powertrain will need to meet many of the normal requirements of the DI gasoline engine. For a four-stroke stoichiometric HCCI/SI powertrain, future legislated emission levels would be met with relatively simple TWC emissions control technology. The low NOx content of the exhaust stream could present some opportunities for reducing catalyst precious metal loadings, diminishing the overall powertrain cost. These NOx levels should also allow some lean HCCI excursions whilst still meeting emission targets by three-way catalysis, resulting in further improved fuel economy. For a two-stroke HCCI-SI powertrain, the lean exhaust stream is likely to dictate the use of lean NOx aftertreatment [17] to meet strict limits for NOx emissions. CYLINDER PRESSURE FEEDBACK
states. In the HCCI engine, the absence of a direct link between a parameter such as spark timing and the combustion event will present a challenge to the conventional EMS. The application of model-based engine control, in particular Cylinder Pressure based Engine Management Systems (CPEMS) may prove to be fundamental in the full realisation of HCCI powertrains [18]. CPEMS employs measurement of cylinder pressure – perhaps the ideal indicator of engine performance – along with model-based control of individual cylinders [19]. CPEMS technology offers the improved adaptive capabilities that are likely to be required in transient control of an HCCI engine. This technology has been enabled by the realisation of inexpensive pressure sensing devices [20]. The combination of CPEMS and HCCI is the subject of a current investigation by Ricardo. TWO-STROKE/ FOUR-STROKE SWITCHING ENGINES A clear trend in engine development has been the pursuit of ever-increasing levels of flexibility in order to meet requirements for fuel consumption, emissions and performance. The ultimate step in flexibility might be the development of engines that can switch between twostroke cycle and four-stroke cycle operation. By the application of a poppet valve combustion system with capability for both two-stroke and four-stroke combustion, such as that described in this paper, and a flexible valvetrain, the optimum combination of four combustion regimes could be selected:
• • • •
Four-stroke SI Four-stroke HCCI Two-stroke SI Two-stroke HCCI
Figure 41 shows the operating ranges for two-stroke HCCI operation and NA four-stroke operation plotted as torque for a 1.4L engine. 140
120
100
Torque [Nm]
It has been shown that the negative overlap approach incurs a reduced, but still significant, pumping work penalty. The total flexibility offered by electromagnetic or electrohydraulic control of the valves should yield further reduction in pumping losses, and camless valve actuation could therefore facilitate a second-generation HCCI engine. Camless valve actuation may also yield larger operating ranges for HCCI.
80
60
20
Expected twostroke HCCI
0 500
Conventional engine management systems are primarily parameter-based systems in which a large number of engine maps are used to control the engine in its many
Two-stroke HCCI Four-stroke HCCI
40
1000
1500
2000
2500
3000
3500
4000
4500
5000
5500
6000
Engine speed [rpm]
Figure 41. Two-stroke and four-stroke HCCI operating regions for a 1.4L engine
Expected two-stroke HCCI operation above 3250 rev/min is shown. Also shown is the published torque curve for a typical four-valve 1.4L engine. The benefit of two-stroke operation in terms of expanded HCCI operation is evident, and the two-stroke results may indeed be the highest specific load demonstrated in an HCCI engine suitable for passenger cars. Two-stroke operation also offers exceptional torque at low engine speeds, making it attractive for downsized engine concepts. Two-stroke/ four-stroke switching engines are the subject of a current investigation led by Ricardo, and with partners including Ford Motor Company, as part of the UK Foresight Vehicle initiative [21].
CONCLUSIONS • HCCI combustion has been demonstrated experimentally in poppet valve DI gasoline engines, using both two-stroke and four-stroke cycle operation • Standard gasoline fuel, a conventional compression ratio, and ambient temperature intake air were used throughout • HCCI operation produced a 99% reduction in NOx emissions and an 8% reduction in ISFC compared with the baseline DI gasoline engine condition for a standard key point • HC emissions for HCCI operation were comparable to other DI gasoline engine modes • The VECTIS CFD code has been successfully used to simulate two-stroke HCCI and SI combustion, providing information on in-cylinder processes including scavenging, spray development and evaporation, ignition and combustion
Begg, Paul Alexander, David Parmenter and Ralph Wood at the University of Brighton for their contributions to the work.
REFERENCES 1. CHRISTENSEN, M.; JOHANSSON, B.; EINEWALL, P. Homogeneous Charge Compression Ignition (HCCI) Using Isooctane, Ethanol and Natural Gas – A Comparison with Spark Ignition Operation SAE 972874, 1997 2. CHRISTENSEN, M.; JOHANSSON, B. Influence of Mixture Quality on Homogeneous Charge Compression Ignition SAE 982454, San Francisco, USA, 1998 3. LAVY, J.; DABADIE, J.; DURET, P.; ANGELBERGER, C.; LE COZ, J-F.; CHEREL, J. Controlled Auto-Ignition (CAI): A New Highly Efficient and Near-Zero NOx Emissions Combustion Process for Gasoline Engine Application A New Generation of Combustion Processes for the Future?, Editions Technip. Paris, France, 2001 4. OAKLEY, A.; ZHAO, H.; LADOMMATOS, N.; MA, T. Experimental Studies on Controlled Auto-Ignition (CAI) Combustion of Gasoline in a 4-Stroke Engine SAE 2001-01-1030, 2001 5. LAVY, J.; DABADIE, J.; ANGELBERGER, C.; DURET, P.; WILLAND, J.; JURETZKA, A.; SCHÄFLEIN, J.; MA, T.; LENDRESSE, Y.; SATRE, A.; SCHULZ, C.; KRÄMER, H.; ZHAO, H.; DAMIANO, L.; Innovative Ultra-Low NOx Controlled Auto-Ignition Combustion Process for Gasoline Engines – The 4-SPACE Project SAE 2000-01-1837, 2000 6. HUNDLEBY, G. Development of a Poppet-Valved Two-Stroke Engine – The Flagship Concept SAE 900802, Detroit, USA, 1990
• The in-cylinder conditions provided by the CFD simulation, such as the in-cylinder temperature and residual distribution, have been used to understand the criteria determining HCCI combustion
7. STOKES, J.; HUNDLEBY, G.; LAKE, T.; CHRISTIE, M. Development Experience of a Poppet-Valved Two-Stroke Flagship Engine SAE 920778, Detroit, USA 1992
• The market introduction of gasoline HCCI has been considered, with appropriate mixture preparation, valvetrain and emissions control technology
8. REITZ, R. D.; DIWAKAR, R. Effect of Drop Breakup on Fuel Sprays SAE 860469, 1986
• The concept of the two-stroke/ four-stroke switching HCCI engine is introduced, enabling significantly expanded HCCI operation
ACKNOWLEDGEMENTS The authors would like to thank the directors of Ricardo plc for their support and permission to publish this paper. Thanks are also extended to Martin Gold, Richard Murphy, Nick Luard, Francis Shields, Chris Wright and the VECTIS development team at Ricardo, and to Steve
9. LIVENGOOD, J. C., WU, P. C.. Correlation of Autoignition Phenomena in Internal Combustion Engines and Rapid Compression Machines 5th International Symposium on Combustion, p.347 1955 10. CHEN, C.; BARDSLEY, M. E. A.; JOHNS, R. J. R. Two-Zone Flamelet Combustion Model SAE 2000-01-2810, 2000
11. Ricardo Software VECTIS Computational Fluid Dynamics Release 3.6 Theory Manual April 2002
CI
Compression Ignition
CO
Carbon Monoxide
COV
Coefficient Of Variation
12. FLIERL, R.; KLÜTING, M. The Third Generation of Valvetrains – New Fully Variable Valvetrains for Throttle-Free Load Control SAE 2000-01-1227, Detroit, USA, 2000
CPEMS
Cylinder Pressure based Engine Management System
DDM
Discrete Droplet Method
DI
Direct Injection
EGR
Exhaust Gas Recirculation
EOI
End Of Injection
14. KANEKO, M.; MORIKAWA, K.; ITOH, J.; SAISHU, Y. Study on Homogeneous Charge Compression Ignition Gasoline Engine Paper 1-18 COMODIA 2001, Nagoya, Japan, 2001
EVC
Exhaust Valve Closing
EVO
Exhaust Valve Opening
FIE
Fuel Injection Equipment
15. LI, J.; ZHAO, H.; LADOMMATOS, N.; MA, T. Research and Development of Controlled Auto-ignition (CAI) Combustion in a Four-Stroke Multi-Cylinder Gasoline Engine SAE 2001-01-3608, 2001
HC
Hydrocarbon
HCCI
Homogeneous Charge Compression Ignition
IC
Internal Combustion
IMEP
Indicated Mean Effective Pressure
ISFC
Indicated Specific Fuel Consumption
IVC
Intake Valve Closing
IVO
Intake Valve Opening
LNT
Lean NOx Trap
MFB
Mass Fraction Burned
NA
Naturally Aspirated
NOx
Nitrogen Oxides
RON
Research Octane Number
RTZF
Ricardo Two Zone Flamelet
SI
Spark Ignition
SOI
Start Of Injection
TDC
Top Dead Centre
TWC
Three-Way Catalyst
ULG
Unleaded Gasoline
VVA
Variable Valve Actuation
13. WILLAND, J.; NIEBERDING, R.; VENT, G.; ENDERLE, C. The Knocking Syndrome – Its Cure and Its Potential. SAE 982483, San Francisco, USA, 1998
16. LAW, D.; KEMP, D.; ALLEN, J.; KIRKPATRICK, G.; COPELAND, T. Controlled Combustion in an IC-Engine with a Fully Variable Valve Train SAE 2000-01-0251, 2000 17. GREGORY, D.; MARSHALL, R.; EVES, B.; DEARTH, M.; HEPBURN, J.; BROGAN, M.; SWALLOW M, Evolution of Lean-NOx Traps on PFI and DISI Lean Burn Vehicles SAE 1999-01-3498, Toronto, Canada, 1999 18. OLSSON, J-A.; TUNESTÅL, P.; JOHANSSON, B. Closed-Loop Control of an HCCI Engine SAE 2001-01-1031, 2001 19. MÜLLER, R.; HART, M., TRUSCOTT, T.; NOBLE, A.; KRÖTZ, G.; EICKHOFF, M.; CAVALLONI, C.; GNIELKA, M. Combustion Pressure Based Engine Management System SAE 2000-01-0928, Detroit, USA, 2000 20. TRUSCOTT, T. A Novel Engine Management System Based on Intelligent Control Algorithms and Utilising Innovative Sensor Technology MST News, 2/99, 1999 21. Foresight Vehicle programme
http://www.foresightvehicle.org.uk/
DEFINITIONS, ACRONYMS, ABBREVIATIONS AFR
Air-Fuel Ratio
ATDC
After Top Dead Centre
BTDC
Before Top Dead Centre
CA
Crank Angle
CAI
Controlled Auto-Ignition
CFD
Computational Fluid Dynamics
