Xin Q. Diesel Engine System Design

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600 Diesel engine system design

the pre-compressor intake throttling. However, any throttle valves placed in

any part of the engine air system increase BSFC signicantly, and their use

should be avoided or minimized in engine system design. Although all the air

control valves can achieve a wide range of controls of exhaust temperature

and space velocity by adjusting the air ow rate, only VVA (e.g., early or

late IVC, or cylinder deactivation) can achieve them without BSFC penalty.

Early EVO and the valve events to obtain internal EGR may also increase the

exhaust temperature. The benet of exhaust temperature increase by VVA

is especially important at light loads and during warm-up.

Enhanced engine braking

VVA allows the braking valve timing and the retarding process efciency

to be optimized at each engine speed so that the retarding power of the

compression-release brake can be maximized (Hu et al., 1997a, 1997b).

Integrated air compressor

Camless VVA can be used in a designated cylinder to deliver the compressed

air to charge the air tank for vehicle use. Such an integrated air compressor

may eliminate the external air compressor on the vehicle.

Switching between two-stroke and four-stroke operations for ring and

engine braking

The ability to switch the engine operation from four-stroke to two-stroke is

attractive due to the high power delivered by the two-stroke operation, for

example during vehicle hill-climbing, acceleration, or deceleration. A camless

VVA device can achieve such an advanced feature on the engine.

Internal EGR for emissions control with diesel VVA

Internal EGR refers to the mass of the residue gas from the previous engine

cycle that is retained in the combustion chamber to participate in the combustion

in the subsequent engine cycle. Although engine development practice in

the past decade has proven that cooled external EGR is necessary to meet

the most stringent emissions regulations for modern diesel engines with the

minimum BSFC penalty, internal EGR nds its role as a useful technology

for diesel HCCI combustion in low-load operation.

The use of hot internal EGR in conventional diesel combustion generally

results in the following: decreased volumetric efciency; increased compression

temperature; reduced oxygen concentration and air–fuel ratio; reduced NO

;

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601Advanced diesel valvetrain system design

increased soot, BSFC, and exhaust temperature; and a large cylinder-to-

cylinder variation of internal EGR rate due to manifold wave effects. A

series of change in internal EGR rate forms a trade-off between NO

and

soot, and another trade-off between NO

and BSFC. Increasing the boost

pressure may improve the trade-offs by simultaneously reducing NO

and

BSFC at the same soot level.

Internal EGR is useful for NO

reduction when negative engine delta P

occurs. The effectiveness of internal EGR on NO

and BSFC varies with

different EGR induction methods in terms of VVA valve events, depending

on the maximum available EGR rate, pumping loss, and the heat loss of

the burned gas. There are three internal EGR methods: (1) negative valve

overlap with early EVC to trap the residue gas; (2) intake valve pre-lift

during the exhaust stroke to induct and store the burned gas in the intake port;

and (3) exhaust valve post-lift during the intake stroke to draw the burned

gas from the exhaust manifold into the cylinder with the exhaust pressure

pulses. The third method seems the best for NO

and BSFC in conventional

diesel combustion (Edwards et al., 1998; Millo et al., 2007), while the rst

method gives the highest cylinder charge temperature which is sometimes

desirable for HCCI. Internal EGR methods and their gas exchange processes

are discussed in detail by Desantes et al. (1995), Benajes et al. (1996), and

Pischinger et al. (2000).

Desantes et al. (1995) and Benajes et al. (1996) used an inline six-cylinder

engine with a divided-entry turbine in their simulation. They found that the

internal EGR rate was inuenced more by the engine speed than the load.

Higher engine speed produced higher EGR rate. They also found that the

intake pre-lift event produced lower EGR rate than the exhaust post-lift event

did. They obtained 5–20% EGR rate at different engine speeds and loads

with different valve timing/event. They discovered that the exhaust post-lift

affected swirl, while the intake pre-lift did not. The swirl ratio was greatly

reduced by the high internal EGR rate attained with the exhaust post-lift.

The Miller cycle and wastegating reduction with diesel VVA

In addition to hot internal EGR, the Miller cycle (Miller, 1947; Miller and

Lieberherr, 1957) is another commonly used feature of VVA application

in the turbocharged diesel engine. The Miller cycle has the following

characteristics: (1) shorter effective compression stroke than the expansion

stroke (i.e., the Atkinson cycle); (2) variable intake valve timing and the

resulting variable effective engine compression ratio; and (3) increased boost

pressure via supercharging or turbocharging to compensate for the reduction

in volumetric efciency and air–fuel ratio caused by the short intake valve

event duration. When the engine geometric compression ratio is reduced, the

expansion ratio is also lowered. This results in a reduction in thermodynamic

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602 Diesel engine system design

cycle indicated efciency. The Miller cycle and the Atkinson cycle can lower

the compression ratio without reducing the expansion ratio by advancing

or retarding the intake valve timing in order to improve the thermodynamic

cycle efciency. The Miller cycle can also reduce peak cylinder pressure

and mechanical load via lower effective compression ratio.

When the Miller cycle uses early IVC during the intake stroke before

the BDC rather than late IVC during the compression stroke, the effect of

internal charge cooling is achieved via cylinder expansion toward the BDC.

Internal cooling can effectively reduce the charge temperature at the BDC and

hence the charge temperature at the start of combustion, resulting in lower

cycle temperatures, thermal loads and NO

. Alternatively, internal cooling

allows the start-of-injection timing to be advanced to achieve better BSFC

at the same NO

level. Previous calculations showed that a 6% reduction in

temperature and pressure can be achieved through adiabatic expansion at

the start of the compression stroke if the intake valve is closed at 60° crank

angle before the intake BDC (Ishizuki et al., 1985). Edwards et al. (1998)

showed in their research of the combined Miller cycle and internal EGR that

internal cooling was achieved to reach the same in-cylinder gas temperature

as that given by the cooled external EGR at the Euro-III level (about 8%

EGR), and using the expansion cooling for the Euro-IV EGR level (18%)

also seemed feasible.

The fuel economy advantages of the Miller cycle result from the following

four aspects: (1) greatly reduced engine delta P and pumping loss; (2) increased

thermodynamic cycle indicated efciency due to reduced compression ratio

relative to the unchanged expansion ratio; (3) increased geometric compression

ratio enabled by the reduced effective compression ratio; and (4) internal

charge cooling due to early IVC. The Miller cycle can reduce NO

due to

lower compression temperature, but may increase soot signicantly if the

air–fuel ratio becomes too low, especially at full load. The NO

reduction

potential of the Miller cycle used with internal EGR is limited to the high-

engines such as HD Euro-III or at LD part load. Cooled external EGR

is required for the Miller cycle used in low-NO

engines. There are six

critical parameters involved in designing an effective Miller cycle: IVC

timing, turbine area, turbine wastegate opening, turbocharger efciency,

compressor matching, and engine geometric compression ratio.

The retarded or advanced IVC timing creates a shorter intake valve event

duration and higher intake manifold boost pressure so that engine delta P

is reduced. The effect of engine delta P reduction by the Miller cycle was

illustrated by Ishizuki et al. (1985). Mavinahally et al. (1996) reported that

the engine with 50% greater expansion ratio than the effective compression

ratio produced 3% higher brake thermal efciency if compared at the same

air–fuel ratio and peak cylinder pressure, although the power capability of

the engine was reduced due to the shortage of air–fuel ratio at full load

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603Advanced diesel valvetrain system design

caused by late IVC. Ishizuki et al. (1985) reported the benet of the Miller

cycle on BSFC by illustrating a signicant BSFC reduction around 3.3% in

a large area in the engine speed–load domain. They also reported that there

was no gain and no loss in BSFC under the high-speed conditions, but this

statement is believed misleading and not conclusive for the Miller cycle.

They mentioned that one remarkable improvement by the Miller cycle was

the increase in BMEP at low speeds by using a smaller turbine area.

In fact, it is found that if compared with the conventional engine at the

same brake power and peak cylinder pressure, the Miller cycle can offer

signicantly lower fuel consumption (e.g., 7%) but at the expense of a little

lower air–fuel ratio due to its low intake manifold volumetric efciency.

If compared at the same air–fuel ratio, EGR rate, brake power, and peak

cylinder pressure, the Miller cycle can still offer signicantly lower BSFC,

usually at least 2% and often in the range of 4–5% at rated power for

modern heavy-duty high-EGR diesel engines. The variation of the BSFC

benet is caused by power density (or engine load), EGR level, engine

speed, turbocharger conguration, etc. The BSFC benet of the Miller cycle

working with wastegating reduction or elimination is especially prominent

at high speeds. The Miller cycle can reduce excess air ow and pumping

loss effectively. Although early or late IVC timing can be used to reduce

engine delta P and pumping loss as low as possible, it should be noted that

an insufcient engine delta P causes the problem of EGR driving, especially

at peak torque. Therefore, xed cam timing usually may not be sufcient

to cover the needs of both delta P reduction and EGR driving at both high

speeds (e.g., rated power) and low speeds (e.g., peak torque). The design

challenge in the Miller cycle is to achieve the variable intake closing timing

required at all speeds via VVA.

The turbine area is usually determined by the EGR-driving and/or the

air–fuel ratio requirements at peak torque with the EGR valve fully open.

The turbine wastegate needs to be adjusted together with the IVC timing in

order to reach the desired air–fuel ratio and engine delta P. The turbocharger

efciency can be adjusted in turbocharger design in order to compensate for

the gap in air–fuel ratio without affecting the engine delta P. It should be noted

that the benet of the thermodynamic cycle efciency of the Miller cycle can

be offset by the increased engine delta P that is caused by a smaller turbine

area, while the engine delta P and BSFC are not negatively affected by the

boost pressure increase that is produced by higher turbocharger efciency.

The compressor size needs to be properly adjusted in order to better t the

engine operating point of the Miller cycle at rated power in the high efciency

region on the compressor map. The boost pressure of the Miller cycle can be

much higher than that in the conventional engine, and the compressor ow

rate can be similar or slightly reduced. The high boost pressure is not an

issue since two-stage turbocharging has been widely used in today’s high-

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604 Diesel engine system design

EGR or high-BMEP engines. In fact, compared to the conventional engine,

the Miller cycle offers an advantage in two-stage turbocharging in the way

that it can split its high boost pressure to the two-stage compressors with

a better distribution on the two compressor maps with higher compressor

efciencies.

The engine geometric compression ratio needs to be increased in the

Miller cycle to match with the reduced effective compression ratio that is

caused by early or late IVC timing. The purpose of increasing the geometric

compression ratio is to fully utilize the design limit of peak cylinder pressure

while maximizing the engine thermodynamic cycle efciency.

Both early and late IVC can affect effective engine compression ratio

and volumetric efciency. The ratio is dened as the volumetric ratio of the

cylinder volume at IVC to the cylinder volume at the TDC. Other denitions

of effective pressure ratio also exist, such as a dynamic pressure-based

compression ratio used by He et al. (2008). A common feature between

early and late IVC is that they both produce very small pumping loss, as

illustrated in the p–V diagram by Pischinger et al. (2000). The difference

between early and late IVC in throttleless operation is that in the p–V diagram

early IVC gives decreasing cylinder pressure after IVC toward the BDC

during the intake stroke as a closed thermodynamic system for the cylinder,

and the cylinder pressure returns along the same path after IVC during the

compression stroke. On the other hand, the late IVC gives almost constant

cylinder pressure during the intake stroke and the compression stroke as an

open thermodynamic system from IVO to IVC.

Usually, late IVC produces slightly higher pumping loss than early IVC

because the valve throttling ow loss during the longer valve opening duration

is higher. Late IVC consumes work to induct the gas mass that is immediately

pumped back out of the cylinder. Early IVC reduces the amount of gas

inducted into the cylinder, and thereby reduces the pumping loss associated

with the induction process. As pointed out by Gallo (1992), high irreversibility

peaks in the irreversibility rate prole of the intake or exhaust process are

associated with the backow through the intake and exhaust valves and the

losses during the mixing process of the hot residue gas and the colder air/

gas mixture. Moreover, late IVC at light loads may cause an increase in the

intake manifold gas temperature due to (1) charge heating by the combustion

chamber walls; and (2) inducting the charge back as hot gas. The heating

effect of charge relling causes a reduction in gas density.

On the other hand, there are concerns with early IVC, too. One problem is

the reduction of in-cylinder air motion and turbulent mixing as consequences

of early IVC (Edwards et al., 1998). Moreover, compared to late IVC, the

practicality of early IVC can be an issue in mechanical cam design due

to the constraints of valvetrain dynamics with short valve event duration

at the same maximum valve lift. Certainly, using lower maximum valve

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605Advanced diesel valvetrain system design

lift is feasible for early IVC in order to make the cam design easier. The

difference between early and late IVC is discussed in detail by Anderson et

al. (1998). The effects of early and late IVC on boost pressure and engine

delta P are shown in the simulation data in Fig. 9.12(b) for a heavy-duty

diesel engine.

In addition to the cam phase shifters and camless devices, rotary valves

have been used as a cost-effective means of VVA to achieve early IVC

timing or the Miller cycle in the gasoline engine (Anderson et al., 1998)

and the diesel engine (Ishizuki et al., 1985; Kamo et al., 1998). The rotary

valve is located upstream of the engine intake valve in the intake port or

manifold. The rotary valve is different from the intake throttle in the way

that it provides a timed open and closed event. The optimum cut-off timing

of the rotary valve can be found at each engine speed and load condition.

The relatively simple mechanism of the rotary valve assures good reliability

and durability, compared with the more complicated modications to the

engine valvetrain in other VVA mechanisms. Ishizuki et al. (1985) conducted

a comprehensive analysis on the advantages and disadvantages of the rotary

valve mechanism. They concluded that its inherent disadvantages such as

the dead volume and leakage did not have a signicant adverse effect on

the Miller cycle performance, although its impact on swirl ratio remained

uncertain. Although the rotary valve technology was initiated for non-EGR

gasoline engines, it may become important for modern high-EGR low-

diesel engines. The rotary valve can also be used to achieve cylinder

deactivation. More information on the rotary valve mechanisms can be found

in Sakai et al. (1985).

The Miller cycle can also be used for boost pressure control in turbocharged

engines. The research conducted by Ke and Pucher (1996) is probably the

most important VVA work (although on the gasoline engine) that sheds light

for turbocharged diesel VVA research. They investigated the feasibility of

using early IVC timing to replace turbine wastegate for boost control in order

to prevent over-boosting for engine protection. They showed that under high

speed (5600 rpm) full load conditions the same cylinder pressure and trapped

air mass at the intake BDC could be achieved by either a conventional cam

with IVC timing at 63° after the BDC with wastegating or an early-IVC cam

with IVC at 59° before the BDC without opening the wastegate. The in-

cylinder temperature was approximately 20 K colder at the intake BDC for

the early-IVC cam. Moreover, the pumping loss of the early-IVC case was

lower than that given by the wastegating case due to lower engine delta P.

The Curtil system with diesel VVA

The above discussion on the Miller cycle shows that some advanced

thermodynamic cycles via valve timing variations may suffer the loss of

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606 Diesel engine system design

air–fuel ratio although they can greatly improve pumping loss, engine thermal

efciency and BSFC. The capability of charging fresh air to maintain an

acceptable air–fuel ratio is important for diesel engines. The traditional way

to compensate for the loss of air–fuel ratio without a negative impact on

engine delta P is to increase turbocharger efciency by better aerodynamic

design of the turbine or the compressor. Innovative ways of charging fresh air

by using exible engine valve events and manifold pressure wave dynamics

without a penalty in pumping loss are also important and attractive.

It is worth considering the ‘Curtil’ system (Curtil, 1998), which can

back-charge the cylinder with the fresh air stored in the exhaust port. In the

Curtil system, the fresh air ows from the cylinder into the exhaust port via

a secondary exhaust valve post-lift event near IVC during the intake stroke

when the intake manifold boost pressure is higher than the exhaust manifold

pressure. Then, the fresh air stored in the exhaust port is pushed back by the

exhaust pulses from adjacent cylinders to return into the cylinder when the

exhaust pressure pulse is greater than the intake port pressure. Therefore,

the trapped air mass in the cylinder can be increased to reach a high air–fuel

ratio for soot reduction or allow a higher fueling rate to produce more torque.

The Curtil system can greatly increase the transient torque and reduce smoke.

Fessler and Genova (2004) reported that 50% higher torque was obtained

using the Curtil system at low engine speeds such as 1000–2000 rpm with

acceptable smoke and BSFC penalty.

It should be noted that the Curtil system only works well in conditions

where the exhaust port pressure oscillates around the intake port pressure

during the intake stroke (e.g., typically at high engine speeds and low

loads). The Curtil system and adjusting EVO timing with VVA are two

primary means of increasing the low-speed torque to improve drivability

and transient performance. The Curtil system is also a promising means for

fresh air charging to supplement other advanced thermodynamic cycles to

compensate for the air–fuel ratio shortage. More detailed information on the

Curtil system is provided by Curtil (1998), Israel (1998), and Fessler and

Genova (2004).

The theory of using wastegating reduction or elimination and VVA (i.e.

WR-VVA or WE-VVA systems) to control pumping loss

Modern high-EGR diesel engines with non-VGT turbines (i.e., fixed

geometry turbines) are characterized by high engine delta P or pumping loss

at high speeds (from part load to full load) due to the requirement of using

a small turbine area to drive sufcient EGR at peak torque or low speeds.

Although VGT is probably the best solution at high speeds and high loads

to solve the problem of high engine delta P, there are concerns about the

VGT technology with regard to the cost, durability, turbine efciency loss at

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607Advanced diesel valvetrain system design

large turbine area openings (e.g., at fully open vane position), and product

availability. Moreover, at engine part load even a good VGT may not have

an advantage over the Miller cycle with VVA. Intake VVA with the Miller

cycle, VGT, and low-pressure-loop EGR (or the hybrid EGR system) are

three competing technologies in terms of engine delta P control (reduction)

in the entire engine speed–load domain.

The application of the Miller cycle with VVA on modern high-EGR diesel

engines offers great advantages in pumping loss control for fuel economy and

air system simplication for cost reduction. It can be foreseen that in the future

HCCI, VVA, and VGT will be the three most important technologies for diesel

engines, one for advanced combustion and the other two for advanced air

systems. Since intake VVA (particularly IVC-VVA) is used as an enabler for

HCCI combustion anyway, using the IVC-VVA or the Miller cycle to extend

its functionality to improve the air system will become naturally logical for

product cost saving (e.g., to avoid the expensive VGT).

The conventional rationale for VVA in diesel engines was to use early IVC

to achieve internal charge cooling due to in-cylinder expansion, or to use early/

late IVC to reduce the effective compression ratio to reduce the compression

temperature and NO

. With other emissions control technologies widely used on

diesel engines (e.g., cooled external EGR system, DPF), the focus of using VVA

should be placed on pumping loss reduction and fuel economy improvement. In

particular, the VVA with early or late IVC timing is very effective in reducing

engine delta P by raising the intake manifold boost pressure. IVC-VVA and

VGT are two solutions for the problem of high pumping loss in modern diesel

engines. VGT can be more effective than IVC-VVA but VGT does not offer

the other advantages of VVA described earlier. Note that the Miller cycle does

not specify how to use the wastegate.

The theory of using IVC timing to control engine delta P and pumping loss

for turbocharged diesel engines is illustrated by a comprehensive simulation

shown in Fig. 9.12(c), which includes the parametric sensitivity effects of the

following key design parameters: IVC timing, turbine area, turbine wastegate

opening, turbocharger efciency, and engine geometric compression ratio.

The simulation is conducted at 3400 rpm rated power for a high-power-

density engine at 20 bar BMEP with a single-stage wastegated turbocharger.

The EGR rate can be xed at any constant value (zero here without losing

the validity for illustration). The second legend in the gure represents the

baseline system. The air system capabilities of different systems corresponding

to ve different legends are presented in the domain of ‘air–fuel ratio vs.

engine delta P’, and the domain of ‘air–fuel ratio vs. BSFC’. The ideal design

target point at high engine speeds should be located at low engine delta P

(with EGR valve fully open) and sufciently high air–fuel ratio in such an

air system capability domain. The relationship between different systems and

their potentials to reach the ideal design target point are clearly illustrated

in the gure, especially in the gure of the air system capability domains.

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608 Diesel engine system design

The principles of the modern diesel engine air system design related to

intake VVA and turbine wastegating reduction or elimination are summarized

as follows.

1. An intake-valve-closing VVA system (preferably early IVC), in

conjunction with using minimum or no turbine wastegating, is highly

desirable for reducing engine delta P (by increasing the intake manifold

boost pressure), pumping loss, and in-cylinder cycle temperature in

order to achieve the best BSFC and the lowest coolant heat rejection

for non-VGT engines. The design goal is to minimize engine delta

P at any speeds/loads so that the required EGR rate can be achieved

with the EGR valve always fully open with minimum EGR circuit ow

restriction.

2. The intake manifold volumetric efciency of the IVC-VVA system needs

to be reduced to around 60–70% via valve event duration change with

an increase intake manifold boost pressure at high speeds in order to

reduce the engine delta P to the minimum level to minimize BSFC.

3. For the IVC-VVA engine, a high geometric compression ratio (CR) can

be designed to maximize thermodynamic cycle efciency and further

enhance cold start capability, compared to the wastegate engine.

4. For xed geometry turbines, the turbine effective area is sized at the

peak torque condition based on the requirements of EGR rate and

air–fuel ratio with the (also considering the high altitude conditions)

EGR valve fully open and usually the intake throttle fully open.

5. The IVC timing at peak torque is optimized with the turbine area for

the required air–fuel ratio and engine delta P to drive sufcient EGR

ow. The IVC timing at rated power or high speeds is optimized to

achieve the best trade-off between air–fuel ratio and engine delta P

(shown by the characteristic lines in the air system capability chart in

Fig. 9.12(c)).

6. Compared to the required air–fuel ratio for combustion and emissions,

if the actual air–fuel ratio delivered by the air system is too low with

early or late IVC at the desirable engine delta P with the EGR valve fully

open (e.g., 0.1–0.2 bar engine delta P in Fig. 9.12(c)), the turbocharger

efciency needs to be increased in order to raise the air–fuel ratio

without affecting the engine delta P signicantly. Or, alternatively the

turbine wastegate can be closed along with a variation in IVC timing

in order to raise the air–fuel ratio in the air system capability chart.

7. If the air–fuel ratio is too high with early or late IVC at the desirable

engine delta P with the EGR valve fully open, the turbocharger efciency

needs to be decreased or the wastegate needs to be opened more.

8. With the IVC-VVA, it is possible to delete the turbine wastegate or make

the EGR valve a less expensive on-off type rather than the continuously

adjustable valve, depending on the air–fuel ratio requirement at different

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609Advanced diesel valvetrain system design

speeds/loads. In other words, the IVC-VVA may be used as an EGR

control device by adjusting engine delta P to replace the EGR valve.

9. The IVC-VVA can replace the intake throttle to reduce air–fuel ratio

to a certain extent. Their difference is on the effect on engine delta

P. The air–fuel ratio reduction by VVA is accompanied by a large

reduction in engine delta P, while the intake throttle does not affect

engine delta P signi cantly.

10. VGT can perform better than VVA in terms of the air system capability

of ‘air–fuel ratio vs. engine delta P’. If the required air–fuel ratio is not

demanding, the IVC-VVA can be used to achieve low pumping loss.

11. The WE-IVC-VVA performs better than wastegating in terms of the

capability of ‘air–fuel ratio vs. engine delta P’ because VVA does not

waste the exhaust energy. This advantage is shown by comparing the

curves in the capability-domain plot in Fig. 9.12(c) at the 0.1 bar engine

delta P. It shows that the IVC-VVA system can deliver signi cantly

higher air–fuel ratio than the wastegating system without VVA at the

same engine delta P. The WE-IVC-VVA system also allows an SOI/

SOC advance.

12. The WE-IVC-VVA is able to adjust volumetric ef ciency by raising the

boost pressure and the location of the engine operating points on the

compressor map. Therefore, a re-match of the compressor is required

in order to  t the operating points as close to the high ef ciency

region as possible. This is particularly important for two-stage turbo.

Compressor outlet air temperature control should be paid attention

since the boost pressure can become very high at rated power in the

IVC-VVA engine.

13. All the questions related to the functionalities of IVC timing, turbine

area, wastegate, turbocharger ef ciency, and intake throttle can be

well explained by Fig. 9.12(c) and equation 9.31. For example, for

the IVC-VVA engine, the four unknowns in equation 9.31 can be p



and



EGREG

with turbine area, wastegate opening, volumetric

ef ciency (related to IVC timing), and EGR valve opening (e.g., fully

open) set as known inputs.

9.8 Variable valve actuation (VVA) for diesel

homogeneous charge compression ignition

(HCCI)

9.8.1 Controlled auto-ignition (CAI) and HCCI

combustion

Controlled auto-ignition (CAI) is a lean combustion mode that does not require

the necessity of a lean exhaust gas aftertreatment on the gasoline engine. CAI

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