2005-01-0166

Lean and Rich Premixed Compression Ignition Combustion in a Light-Duty Diesel Engine

Timothy J. Jacobs, Stanislav V. Bohac, Dennis N. Assanis University of Michigan

Patrick G. Szymkowicz GM R&D and Planning

Copyright © 2005 SAE International

ABSTRACT

Lean premixed compression ignition low-temperature combustion promises to simultaneously reduce NOx and PM emissions while suffering a moderate penalty in fuel consumption. Similarly, opportunities exist to develop rich combustion strategies which can provide the necessary exhaust constituents for aggressive regeneration of a lean NOx trap (LNT). The current work highlights the development of lean and rich premixed compression ignition combustion strategies. It is shown that the lean premixed compression ignition combustion strategy successfully operates with low NOx and smoke, at the expense of a 5% increase in fuel consumption over conventional diesel operation. The rich premixed compression ignition combustion strategy similarly operates with low NOx and smoke, and produces enough CO (up to 5% by volume in exhaust) for aggressive regeneration of an LNT.

INTRODUCTION

In his lecture given to the Society of Cassell on June 16, 1897, Rudolph Diesel stipulated as a third condition of his rational heat motor that fuel must be introduced gradually so as to maintain an isothermal combustion process [1]. Hence, from the early days of diesel engine development, it appeared that diffusion burn combustion would dominate. Nevertheless, this developmental road progressively changed direction as awareness of vehicle emissions and their impact on the atmosphere surfaced [2]. As researchers learned more about diesel engine combustion, it became increasingly clear that the diffusion burn portion was largely responsible for its soot emission [3]. Therefore, the desire to overturn Diesel’s condition of isothermal combustion developed, and attention shifted to premixed combustion modes [4].

Today, the development of combustion strategies resembling homogenous charge compression ignition strategies is vigorously pursued. The promise of simultaneously reduced NOx and PM offers attractive

incentives, especially when considering the associated minor penalties in fuel economy. The popular press has become excited at the prospects of HCCI-type combustion systems, which are viewed as the internal combustion engine’s best response to future competition from fuel cells and hybrids [5].

Much of the developmental strategies and targets are dictated by upcoming stringent emissions standards on passenger vehicles. One promising area of development focuses on preparing a premixed charge that burns at low combustion temperatures. Kimura et al. [6] and Akihama et al. [7] have demonstrated this capability by applying heavy levels of EGR to increase combustion ignition delay. The increased ignition delay ensures more complete mixing of the air and fuel prior to combustion so as to avoid hot flame regions. The high EGR levels also maintain low combustion temperatures - temperatures below those required to pyrolize fuel and form the precursors to soot. The added benefit demonstrated by this development is the ability to run the engine rich, offering opportunities to regenerate aftertreatment systems through in-cylinder processes.

From the previously published literature on lean and rich premixed combustion, a need to fundamentally understand such combustion strategies arises in order to tackle the challenges associated with them and deploy them successfully. The present study attempts to provide this insight and identifies the fundamental mechanisms for simultaneous reductions in NOx and PM. Our objectives are to develop lean and rich premixed compression ignition strategies, to explore their merits and drawbacks, and to understand the fundamental differences between such strategies and a conventional diesel strategy.

The study is structured by first laying a fundamental foundation, through the use of classical conceptual models, for understanding conventional diesel combustion. These classical models are extended to the new mode of combustion, where simultaneous reductions of NOx and PM are observed. An eventual

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disconnect between observed trends and expected trends signals the migration to low temperature combustion. With this new opportunity of low temperature combustion, a rich PCI condition demonstrates the possibility of diesel aftertreatment regeneration without high levels of NOx and soot. Finally, the study concludes by comparing the various combustion strategies, and identifying tradeoffs between them.

EXPERIMENTAL SETUP

The experimental development of a premixed compression ignition combustion strategy took place on a production-type 1.7 L diesel engine. The 4-cylinder engine utilizes a common rail fuel system, variable geometry turbocharger, exhaust gas recirculation, and an intake throttle. It has four valves per cylinder, swirl control valves, centrally located injectors, and a bowl-in piston design. The production calibration of this engine is certified to meet Euro IV emission standards, with a rating of 100 horsepower at 4400 RPM [8]. The engine is located in the General Motors Collaborative Research Laboratory at the University of Michigan’s WE Lay Automotive Laboratory.

Air flow rate is measured using a laminar flow element and fuel flow rate is measured using a positive displacement meter. Manifold temperatures and pressures are measured using thermocouples and strain-based pressure transducers, respectively. In-cylinder pressure measurements in all four cylinders are collected from Kistler 6041B water-cooled dynamic pressure transducers. Combustion noise is measured with an AVL 450 Noise Meter. Gaseous exhaust emissions are collected with an AVL CEB II emissions bench, where all sample lines are heated to 190°C. Smoke measurements are made using an AVL 415 filter smokemeter.

Calculations of heat release profiles use pressure data from cylinder 1, and assume constant mass throughout the cycle (i.e., neglected mass changes due to fuel injection, crevice flow, and blow-by). Brunt and Platts [9], along with Krieger and Borman [10], provide correlations for determination of thermodynamic properties, while Hohenberg’s [11] correlation is used for estimation of in-cylinder heat transfer.

Emission targets for development of each combustion strategy are set using emission indices, which are fuel mass specific measures of a particular exhaust species. Similarly, air-fuel ratios are calculated from emission measurements, indicating combustion air-fuel ratio. Due to the oxygen and fuel concentrations in the exhaust gas recirculation, differences exist between combustion air-fuel ratio and engine-induced air-fuel ratio. The emissions index for soot is calculated from smoke measurements using the MIRA correlation [12].

The development objectives of the following combustion strategies consider the inclusion of aftertreatment devices. As such, low-sulfur fuel was used throughout the entire combustion development. Table 1 provides a comparison between this low-sulfur fuel and commercial diesel #2.

The engine was modified in two ways from its true-to-production hardware state. The first modification is the redesign of the EGR cooler. Due to the higher EGR flow, a larger cooler is required to cool the EGR. The cooler was sized such that EGR could be cooled to a temperature close to the engine coolant temperature (90°C). The second alteration is a redesign of the piston bowl to reduce cylinder compression ratio from 19:1 (production compression ratio) to 16:1 [13]. This reduction in compression ratio reduces the pre-injection thermodynamic state, resulting in increased ignition delays. The increased ignition delay increases the time for fuel and air mixing, allowing for attainment of premixed compression ignition combustion.

All combustion development took place using the engine’s electronic control module. Controllable parameters include injection duration, injection pressure, injection timing, boost pressure, air and EGR rate (through the control of the EGR valve, variable geometry turbocharger, and intake throttle position).

OPERATING CONDITIONS AND TARGETS

Three combustion modes were tested in this study. The lean conventional mode represents a typical present-day calibration, while the lean and rich PCI modes represent the new strategies being studied. Rail pressure, EGR rate, and injection timing were optimized for each of the new PCI modes.

Combustion development targets, provided in Table 2, guide the research performed for this study. The emissions targets were determined in consideration of future vehicle emission targets. A single fuel injection event and constant fuelling rate are used for each combustion mode. The one exception is the rich PCI condition, where fuel rate had to be increased to generate rich products and still meet the load

Low – Sulfur Diesel

Diesel #2

Cetane Number 52 50 Sulfur Concentration (ppm) 12 <500 Lower Heat Value (MJ/kg) 43.481 42.91

A/F Stoichiometric 14.74 14.46 Density (kg/m3) 810 840

T50 (K) 530 498 Table 1: Fuel properties of Low-Sulfur Diesel (used in this study) in comparison to current-day production of Diesel #2.

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requirements. All data were collected at steady-state, fully-warmed engine operation.

LEAN CONVENTIONAL COMBUSTION DEVELOPMENT

Combustion development for this research study starts with understanding a baseline conventional combustion mode. The premixed compression ignition combustion described in the following sections is compared to the baseline conventional combustion strategy described in this section. This comparison offers insights as to how the premixed strategies simultaneously reduce soot and NOx. The conventional strategy can also serve as a suitable lean NOx trap loading condition, where NOx values are relatively higher, yet fuel consumption is relatively lower. An attempt to correlate the observed trends to classical understanding of diesel combustion provides support for explaining emission-reducing trends.

The effect of injection timing on NOx and smoke is shown in Figure 1. NOx decreases monotonically as timing is retarded. Initially, smoke increases as expected based on classical diesel combustion behavior. However, as injection timing is further retarded, smoke begins to drop.

To explain the NOx behavior, one considers its dependence on flame temperature and residence time. In spite of EGR rate remaining relatively constant throughout the injection timing sweep, flame temperatures decrease as injection timing is retarded [14].

Smoke (or soot) in the exhaust is the difference between soot formation and soot oxidation [15]. Soot formation occurs as fuel pyrolizes during high temperature combustion. Fuel pyrolysis is defined as hydrocarbon chain fragmentation in the absence of oxygen. These fragmentations then develop into nucleation sites for

hydrocarbons and sulfates to adhere onto, thus forming the soot particle. Soot oxidation occurs as high temperature gases promote soot burning. As Khan et al. [15] suggest, and Figure 1 and Figure 2 confirm, the injection timing resulting in the shortest ignition delay yields the highest level of smoke. Based on Lyn’s conceptualization of diesel combustion [16], a shorter ignition delay yields a higher percentage of diffusion burn because less time is available for mixing prior to ignition. As expected, combustion duration is maximized at the same injection timing that ignition delay is minimized. Since diffusion combustion has a slower burn rate than premixed combustion, it substantiates the claim that combustion duration increases as the diffusion burn portion increases. Therefore, smoke increases as the contribution of diffusion burn to the total heat release increases.

Figure 2: Ignition delay and combustion duration versus injection timing for the lean conventional condition at 1500 RPM, 3.75 bar BMEP, 300 bar rail pressure, 32% EGR rate, variable injection timing.

Figure 1: EI-PM and EI-NOx versus injection timing for the lean conventional condition at 1500 RPM, 3.75 bar BMEP, 300 bar rail pressure, 32% EGR rate, variable injection timing.

Lean Conventional

Lean PCI

Rich PCI

Speed (RPM) 1500 1500 1500 BMEP (bar) 3.5 to 4 3.5 to 4 3.5 to 4

Rail Pressure (bar) 300 1000 1000 EGR Rate (%) 31 to 35 41 to 45 41 to 50

Inj Timing (°BTDC) 3 to 15 9 to 18 7 to 26 EI-NOx (g/kg-fuel) <4.5 <1.0 <0.5 EI-PM (g/kg-fuel) <0.5 <0.1 <0.1

Noise (dB) <90 <90 <90 Fuel Penalty (%) Base <5 --

CO Level (%) -- -- >5

Table 2: Combustion development modes and the corresponding operating conditions and development targets.

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Khan et al. [15] suggest that in order to reduce net soot release, one must decrease its rate of formation by decreasing the proportion of diffusion burn. Others have argued that it’s more effective to increase soot oxidation to reduce net soot release [17]; however it’s clear by observing Figure 1 and Figure 2 that a direct relationship exists between smoke and combustion duration; or, diffusion burn where soot is formed.

A desirable lean conventional combustion strategy can be chosen from the injection timing sweep. Figure 3 shows that two conditions satisfy the emission and noise targets listed in Table 2. Of these two conditions, Figure 4 identifies 5° BTDC as the preferred timing to minimize BSFC while achieving the PM and NOx targets. The lean conventional combustion strategy has thus been chosen; the final combustion strategy and emissions are shown in Table 3. A lean premixed compression ignition strategy that further reduces NOx and PM now follows.

LEAN PREMIXED COMPRESSION IGNITION COMBUSTION DEVELOPMENT

In spite of the success of meeting the emission targets with a conventional combustion strategy, even more stringent future emission standards may require novel modes of combustion. One such mode, premixed compression ignition (PCI), attempts to maintain fully premixed combustion during the entire fuel energy release period.

There are multiple ways of creating PCI combustion; however, doing so without EGR raises combustion noise and in-cylinder pressures to unmanageable levels. Therefore, for the following PCI strategy EGR rate will be increased above that used for a conventional combustion strategy. When PCI combustion has been attained, smoke levels should drop due to the elimination of diffusion burning, as suggested by Khan [15]. Furthermore, due to the increased levels of EGR, combustion temperatures drop thereby also lowering NOx formation. By employing a PCI strategy with high EGR rates, one observes simultaneous reductions in PM and NOx.

Increasing the EGR rate increases the ignition delay, allowing more time for air-fuel mixing. The lowering of the compression ratio from 19:1 to 16:1 also assists with increasing ignition delay. Additionally, advanced timings and higher rail pressures ensure complete fuel injection and mixing prior to ignition. While these control parameters assist with attaining fully premixed combustion, the increased EGR also promotes the possibility of low temperature combustion [6]. As will be shown, achievement of low temperature combustion in conjunction with PCI can have profound benefits on PM and NOx reduction.

The impact of EGR and injection timing on EI-NOx is shown in Figure 5. Notice that EGR rates of 42 and 43% cannot meet the development target of EI-NOx < 1 g/kg-fuel for early injection timings. However, the

Figure 4: Brake specific fuel consumption and EI-PM versus EI-NOx, for the lean conventional condition at 1500 RPM, 3.75 bar BMEP, 300 bar rail pressure, 32% EGR, variable injection timing.

Lean Conventional

Speed (RPM) 1500 BMEP (bar) 3.96

Rail Pressure (bar) 300 EGR Rate (%) 32

Inj Timing (°BTDC) 5 EI-NOx (g/kg-fuel) 4.28 EI-PM (g/kg-fuel) 0.34

Noise (dB) 88.2 BSFC (g/kW-hr) 235

Table 3: Operating condition and emissions for the lean conventional combustion mode.

Figure 3: EI-PM and Noise versus EI-NOx for the lean conventional condition at 1500 RPM, 3.75 bar BMEP, 300 bar rail pressure, 32% EGR, variable injection timing.

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development target is met at each timing for EGR rates of 44 and 45%.

The prospects of PCI combustion are elucidated by comparing the results of Figure 5 to those of Figure 6. For all timings and EGR rates, the EI-PM level is below the combustion development target of 0.1 g/kg-fuel. The behavior of EI-PM for the lean PCI condition resembles the behavior for that of the lean conventional condition (Figure 1). However, Figure 6 identifies a unique departure at the higher EGR rates of 44 and 45%. At these EGR rates, as timing is retarded, EI-PM drops more rapidly than at the milder EGR cases.

Unfortunately, one cannot rely simply on ignition delay behavior to explain this observation. Khan et al.’s [15] observation that the minimum ignition delay creates the highest amount of smoke does not apply to this strategy, largely because it burns as predominantly premixed

combustion. Figure 7 supports this statement, as some of the shortest ignition delays create the lowest levels of smoke (see 42% EGR), while some of the longest ignition delays create the highest levels of smoke (see 45% EGR). Of course, EGR plays the driving role with these two examples by cooling combustion and reducing post-flame oxidation of previously formed soot [18].

The unique behavior of smoke versus injection timing (and EGR rate) indicates the transition into a new mode of combustion: low temperature combustion. The impact of temperature on soot formation was extended beyond Khan’s observations by Ahmad, et al. [19] Ahmad’s study demonstrated that carbon formation rate is primarily influenced by temperature. However, high temperature also assists carbon oxidation, which similarly depends strongly on temperature.

Kamimoto and Bae [20] summarize this temperature effect, as well as the role that localized equivalence ratio plays in this scenario. An inverted parabola characterizes soot formation dependency on temperature. At low temperatures (1600 K), soot formation ceases regardless of the local equivalence ratio as temperatures are too low for fuel pyrolysis. At high temperatures (2400 K), the oxidation of hydrocarbon fragmentations occurs faster than their development into nucleation sites, thus lowering the net soot yield.

Within this temperature band (between 1600 K and 2400 K), a lower A/F yields higher soot formation. Thus, the support of Khan’s observations that lower levels of diffusion burn yield less soot is borne from lowering local equivalence ratios that decrease with increasing ignition delay. While Kamimoto and Bae suggest using high combustion temperatures to reduce diesel engine smoke, other researchers suggest migrating to the other extreme of low temperature combustion to mitigate the formation of smoke all together [6] and [7].

Figure 6: EI-PM versus injection timing for the lean PCI condition at 1500 RPM, 3.75 bar BMEP, 1000 bar rail pressure, variable EGR rate, variable injection timing.

Figure 7: Ignition delay versus injection timing for the lean PCI condition at 1500 RPM, 3.75 bar BMEP, 1000 bar rail pressure, variable EGR rate, variable injection timing.

Figure 5: EI-NOx versus injection timing for the lean PCI condition at 1500 RPM, 3.75 bar BMEP, 1000 bar rail pressure, variable EGR rate, variable injection timing.

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In addition to the reduction of locally rich zones through improved premixing, this region of low temperature combustion explains the behavior observed at the higher EGR rates and late timings for the lean PCI combustion strategy. Figure 8, which illustrates EI-PM versus EGR rate at an injection timing of 9° BTDC, shows that an increase in EGR rate from 42% to 43% results in a higher soot yield (marked by line A-B). The increase in soot likely results from decreased oxidation as flame temperature decreases with EGR. Although flame temperature decreased, it is still high enough that fuel pyrolysis, and thus soot formation, occurs.

As EGR rate increases from 43% to 44% (line B-C), flame temperatures begin to fall below the fuel pyrolysis temperature. Therefore, while soot oxidation slows, soot formation slows at a faster rate with the increased level of EGR. Clearly, as EGR rate increases from 44% to 45% (line C-D), soot formation begins to cease entirely as net soot release falls below the level of 42% EGR.

Also shown in Figure 8 is the EI-PM dependence on EGR rate at 15° BTDC injection timing. Notice for this timing, unlike the 9° BTDC timing, that EI-PM continues to increase as EGR rate increases. For this particular timing, the combustion mode never switched completely into low temperature combustion.

Supporting these claims of low temperature combustion, Figure 9 shows the rates of heat release at the conditions presented in Figure 8. At the lower EGR rates of 42 and 43%, the heat release curves are very similar and show a relatively high rate of heat release. As EGR rate increases to 44%, the rate of heat release diminishes, and shifts to later in the cycle. Finally, increasing the EGR rate to 45% dramatically decreases the rate of heat release and further shifts it to later in the cycle. This reduction in the rate of heat release, as well as the shifting of the heat release to the expansion stroke, results in cooler combustion.

The data shown in Figure 10 confirm the shift into low temperature combustion. This illustration, providing the rates of heat release at various timings and an EGR rate of 45%, suggests that as timing is retarded, combustion temperatures decrease dramatically. Notice that the rate of heat release shifts by much more than 3 degrees as timing is retarded from 12° BTDC to 9° BTDC. This shift, as well as the decreased rate of heat release, results in combustion temperatures low enough to avoid fuel pyrolysis, and thus soot formation.

A depiction of both noise and brake specific fuel consumption closes this discussion of the lean PCI combustion development. Notice from Figure 11 that there are few conditions that satisfy the noise constraint of 90 dB. Effectively, EGR rate must exceed 43%, and injection timing cannot advance beyond 15° BTDC for this particular engine and operating condition. Fortunately, after evaluation of Figure 5, there is good

Figure 9: Rate of heat release versus engine crank angle for the lean PCI condition at 1500 RPM, 3.75 bar BMEP, 1000 bar rail pressure, variable EGR rate, 9° BTDC injection timing.

Figure 8: EI-PM versus EGR Rate for the lean PCI condition at 1500 RPM, 3.75 bar BMEP, 1000 bar rail pressure, variable EGR rate, 9° and 15° BTDC injection timings.

Figure 10: Rate of heat release versus engine crank angle for the lean PCI condition at 1500 RPM, 3.75 bar BMEP, 1000 bar rail pressure, 45% EGR rate, variable injection timing.

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alignment between strategies that meet the NOx development target and the noise constraint. Therefore, the choice of a final lean PCI strategy depends on the optimization of fuel consumption.

Brake specific fuel consumption at various EGR levels and injection timings is shown in Figure 12. Working within the constraints of the combustion development targets, the optimal fuel consumption condition occurs at an EGR rate of 45% and an injection timing of 15° BTDC. While an injection timing of 18° yields a better BSFC than 15°, its noise level exceeds the development target of 90 dB.

There is an interesting trend of a downward BSFC curve for increasing EGR rate in Figure 12. This trend also supports the notion of low temperature combustion; as combustion temperatures cool, heat transfer through the cylinder walls decreases. Thus, less energy is lost

through heat transfer, and fuel consumption improves even as EGR rate increases.

When determining the optimal lean PCI condition, there is the option to chose the 9° timing. This timing offers lower EI-NOx and EI-PM relative to the 15° timing. However, this timing was not chosen since both EI-NOx and EI-PM are within the development targets at 15° BTDC injection timing and fuel consumption is better. Therefore, the use of low temperature combustion is not needed to meet the development targets; however it is available if lower EI-NOx and EI-PM are desired. The chosen strategy for lean PCI combustion is summarized in Table 4.

RICH PREMIXED COMPRESSION IGNITION COMBUSTION DEVELOPMENT

With the demonstration of low temperature combustion at the lean condition, one recognizes that there no longer exists a dependence of air-fuel ratio on soot formation. With this in mind, the possibility of running the engine rich for aftertreatment regeneration exists [7]. Rich PCI combustion has the potential of very low NOx and soot emissions, but it is not practical for steady-state operation because of high fuel consumption.

Therefore, development will stay focused to achieving a suitable rich regeneration strategy. Such a strategy could work effectively with a lean NOx trap (LNT). After a certain amount of time, the storage (or loading) efficiency of the LNT declines. Regeneration of the trap requires introducing the device with exhaust constituents that favorably desorb the stored molecules and reduce them to molecular nitrogen [21].

The following describes an aggressive LNT regenerative combustion strategy, where the engine produces upward to 5% carbon monoxide. Because the combustion strategy has entered the low temperature combustion regime, both NOx and PM are very low. This feature is important, so that soot contamination and NOx interference are non-issues during a rich regeneration.

Figure 12: Brake specific fuel consumption versus injection timing for the lean PCI condition at 1500 RPM, 3.75 bar BMEP, 1000 bar rail pressure, variable EGR rate, variable injection timing.

Figure 11: Noise versus injection timing for the lean PCI condition at 1500 RPM, 3.75 bar BMEP, 1000 bar rail pressure, variable EGR rate, variable injection timing.

Lean Conventional

Lean PCI

Speed (RPM) 1500 1500 BMEP (bar) 3.96 3.75

Rail Pressure (bar) 300 1000 EGR Rate (%) 32 45

Inj Timing (°BTDC) 5 15 EI-NOx (g/kg-fuel) 4.28 0.31 EI-PM (g/kg-fuel) 0.34 0.07

Noise (dB) 88.2 88.7 BSFC (g/kW-hr) 235 246

Table 4: Operating condition and emissions for the lean PCI combustion mode.

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Since the regeneration strategy attempts to efficiently maximize carbon monoxide, Figure 13 illustrates the relationship between CO and exhaust air-fuel ratio. As air-fuel ratio decreases below stoichiometry, CO and HC rise dramatically. In general, maximizing CO more effectively regenerates an LNT due to the water-gas shift reaction [22].

Figure 14 illustrates the tradeoff between fuel consumption and carbon monoxide production. A more aggressive regeneration (i.e. higher concentration of carbon monoxide) consumes more fuel. Of course, a more aggressive regeneration requires less operating time. Therefore, proper determination of the regenerative strategy requires optimization.

Another constraining parameter to consider is combustion noise. Figure 15 demonstrates that at less aggressive regenerative conditions (between 3 and 5%

CO), noise levels exceed the developmental target of 90 dB. Therefore, a mild regenerative strategy generates too much noise, preventing its use in spite of better fuel consumption. The physical reason for the noise trend, as well as data scatter, is the method in which the various air-fuel ratios were created (variation of both injection timing and EGR rate).

In summary, an aggressive regenerative strategy was chosen to represent the rich PCI condition. In spite of the higher fuel consumption, the aggressive condition generates more CO for rapid LNT regeneration, reduces the amount of time the engine runs rich, and maintains noise levels below 90 dB. Table 5 summarizes the operating condition and emissions of the rich PCI combustion mode.

COMPARISON OF COMBUSTION STRATEGIES

The following discussion summarizes and compares the combustion strategies described in the previous

Figure 13: Carbon monoxide and total hydrocarbons versus air-fuel ratio for rich PCI condition at 1500 RPM, 3.75 bar BMEP, 1000 bar rail pressure, variable EGR rate, variable injection timing.

Figure 14: Brake specific fuel consumption versus carbon monoxide concentration for the rich PCI condition at 1500 RPM, 3.75 bar BMEP, 1000 bar rail pressure, variable EGR rate, variable injection timing.

Lean Conventional

Lean PCI

Rich PCI

Speed (RPM) 1500 1500 1500 BMEP (bar) 3.96 3.75 3.66

Rail Pressure (bar) 300 1000 1000 EGR Rate (%) 32 45 50

Inj Timing (°BTDC) 5 15 25 EI-NOx (g/kg-fuel) 4.28 0.31 0.05 EI-PM (g/kg-fuel) 0.34 0.07 0.01

Noise (dB) 88.2 88.7 88.8 BSFC (g/kW-hr) 235 246 307

CO (%) -- -- 5.56

Table 5: Operating condition and emissions for the rich PCI combustion mode.

Figure 15: Noise versus carbon monoxide concentrations for the rich PCI condition at 1500 RPM, 3.75 bar BMEP, 1000 bar rail pressure, variable EGR rate, variable injection timing.

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sections. Clearly, one of the most important aspects of novel combustion development is fuel consumption. As such, a “best BSFC” point was generated, indicating the engine’s best fuel economy at the given load condition with optimal injection timing, rail pressure, boost pressure, and no emission constraints. The various fuel consumption levels are shown in Figure 16. Notice that the fuel consumption penalty between the lean PCI condition and the lean conventional condition is 5%. The same comparison to the best BSFC condition yields a 10% fuel consumption penalty.

However, these penalties come with considerable benefits, as shown in Figure 17 and Figure 18. The emission of NOx is reduced by 93% when running a lean PCI condition as opposed to a lean conventional condition. Similarly, the emission of EI-PM is reduced by 79%. Notice from Figure 18 the increase in EI-PM between the best BSFC condition and the lean conventional condition. This increase results from the lower combustion temperatures caused by EGR and decreased excess oxygen. However, the EI-PM decreases between the lean conventional condition and the lean PCI condition (and slightly from the best BSFC point). This decrease in PM indicates the transition to low temperature combustion. The low PM associated with the best BSFC condition results from high temperature combustion (and hence high NOx).

If the air-fuel ratio is further reduced to the rich PCI condition, NOx and smoke virtually disappear. As more fuel is added to the cylinder, assuming that EGR, injection timing, and rail pressure are adjusted to avoid premature combustion and thus localized hot zones, there is a combustion cooling effect that takes place. This combustion cooling results in even cooler combustion temperatures, thus significantly reducing NOx and PM. Of course, fuel consumption at this condition prevents application of this strategy on a continuous basis in a vehicle.

The air-fuel ratio at the various conditions, shown in Figure 19, steadily decreases with the application of higher EGR. The flow of EGR at three of the conditions increases the engine’s pumping losses, especially when a throttle is employed. Therefore, in spite of a higher energy mixture, the engine power decreases due to decreasing volumetric efficiency and fuel conversion efficiency. Figure 20, illustrating the engine’s brake mean effective pressure, quantifies this downward trend.

Another area of concern surrounding the development of a PCI strategy is the rise in carbon monoxide and hydrocarbons. A decrease in combustion temperatures results in decreased post flame oxidation for these molecules. As Figure 21 and Figure 22 illustrate, the lean PCI condition shows a considerable rise in CO and HC relative to a lean conventional strategy. Of course for the rich PCI condition CO and HC emissions rise dramatically; this being the primary objective of such a strategy.

A promising area of technology that can assist with the high emission of CO and HC is the diesel oxidation catalyst (DOC). As any catalyst, the DOC requires a minimum gas temperature to oxidize the respective species. Premixed compression ignition conditions have considerably lower temperatures than lean conventional conditions, identified by Figure 23. This, along with high CO and HC emissions, poses challenges for PCI combustion implementation and will be addressed in future studies.

Figure 16: Brake specific fuel consumption at various combustion development conditions. Best BSFC condition optimizes power (at similar fuelling levels) with timing, rail pressure, boost pressure, and no

Figure 17: EI-NOx at various combustion development conditions. The best BSFC condition operated without any emission control.

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Figure 18: EI-PM at various combustion development conditions.

Figure 23: Exhaust temperature for the various combustion development conditions.

Figure 20: Brake mean effective pressure for the various combustion development conditions.

Figure 22: EI-HC for the various combustion development conditions.

Figure 21: EI-CO for the various combustion development conditions.

Figure 19: Air-Fuel Ratio at various combustion development conditions.

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SUMMARY AND CONCLUSIONS

In an attempt to improve diesel engine NOx and smoke emissions, this study has assessed the merits of both lean and rich premixed combustion strategies.

A lean premixed compression ignition low-temperature combustion strategy was successfully developed using high rates of exhaust gas recirculation, high fuel injection pressures, and injection timings near TDC. When compared with a classical conventional diesel combustion pattern, the new strategy demonstrates the capability to significantly and simultaneously reduce both PM and NOx emissions. Furthermore, the novel strategy exhibited behavior that deviated from classical combustion. This deviation indicated the transition of combustion to low temperature regions where soot particle precursors fail to develop. As a result, the strategy reduced NOx by 97%, and smoke by 79% relative to the lean conventional strategy. Nevertheless, this was accomplished at the expense of a 5% penalty in fuel consumption, as well as higher CO and hydrocarbons emissions. While a diesel oxidation catalyst could assist with the removal of these species, the accompanying low exhaust temperatures may hinder its effective use.

A rich PCI condition that can aggressively regenerate a lean NOx trap accompanied the combustion development efforts. Aside from producing high levels of CO and HC for regeneration, the rich PCI condition offers extremely low engine emissions of nitric oxides and smoke. However, the considerably higher fuel consumption, carbon monoxide, and hydrocarbon levels likely prevent rich PCI combustion as a standard engine operating calibration. The benefit of rich PCI yielding low smoke and NOx becomes apparent when considering the regeneration of an LNT. Low soot engine emission avoids the propensity to mask aftertreatment systems with soot. Low NOx engine emission prevents its interference with the NO2 undergoing reduction in the catalyst.

ACKNOWLEDGEMENTS

The authors of this research study would like to thank the technical and financial resources of the General Motors Collaborative Research Laboratory. The GM CRL provides the focal point for joint research between General Motors Research and Development and the University of Michigan, establishing the framework to link faculty and student expertise with long-term GM needs and to motivate the growth and strengthening of additional areas of excellence of importance to GM. Roger Krieger and John Pinson are thanked for their technical insights and direction in planning and developing this research work. Similarly, University of Michigan machine shop personnel Bill Kirkpatrick and John Mears are thanked for their contribution to the engine test cell setup.

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