Xin Q. Diesel Engine System Design

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610 Diesel engine system design

engines usually use VVA to achieve the controls on air charge mass, gas

composition and charge temperature in addition to the benet of pumping

loss reduction due to the throttleless operation. CAI signicantly improves

fuel economy and greatly reduces NO

. CAI research was summarized by

Zhao (2007).

Homogeneous charge compression ignition (HCCI) is the counterpart of

CAI for diesel and other compression ignition engines. HCCI is a hybrid

strategy between the traditional homogeneous charge spark ignition (used

in the conventional gasoline engine) and the stratied charge compression

ignition (used in the conventional diesel engine). It refers to the auto-ignited

combustion of premixed fuel–air–diluent mixture under lean and/or diluted

conditions with burned gas. The spontaneous ignition occurs simultaneously

at many sites throughout the combustion chamber with an optimum auto-

ignition timing close to the TDC in order to maximize the indicated efciency.

Fuel consumption can be improved by HCCI due to its short and efcient

heat release rate. The absence of a high-temperature ame front leads to

practically negligible formation of nitrogen oxides, and the absence of fuel-rich

zones in the homogeneous lean mixture leads to very little soot formation,

thus resulting in a simultaneous great reduction in both NO

and particulate

matter. HC and CO emissions of HCCI are usually high. Sometimes they

can be at the same levels as the direct injection gasoline engine. The primary

objective of diesel HCCI is to reduce NO

and soot so that the use of NO

aftertreatment can be avoided. HCCI research is summarized by Zhao et al.

(2003) and Zhao (2007).

As pointed out by Pastor et al. (2007) in their review about conventional

and HCCI diesel combustion, how to create a homogeneous mixture, how

to ignite such a mixture, and how to control the combustion are three major

aspects of HCCI technology. The challenges in HCCI include ignition timing

and combustion phasing controls, restricted engine operating load range,

and high HC and CO emissions. Unlike the conventional diesel combustion

where the start of combustion and combustion phasing are controlled by fuel

injection characteristics, the start of combustion and the heat release rate of

HCCI combustion cannot be controlled by fuel injection rate. They depend

only on the reaction kinetics of the cylinder charge. HCCI ignition is controlled

by the cylinder charge temperature during the compression, which is affected

by the following three factors: (1) initial temperature and composition (e.g.,

EGR or residue gas) of the intake charge mixture; (2) the rate of compression

(i.e., temperature evolution along the crank angle during the compression); and

(3) the extent of compression, which is affected by the effective compression

ratio. Note that compared to the gasoline engine, the diesel engine has a higher

geometric compression ratio and the diesel fuel has a lower ignition temperature

due to its lower octane number. Therefore, the HCCI diesel engine relies on

intake charge heating and hot burned exhaust gas to trigger auto-ignition to a

lesser extent than the CAI gasoline engine does.

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611Advanced diesel valvetrain system design

There are different methods of fuel injection to form the mixture in

diesel HCCI, such as port injection and in-cylinder fuel injection during the

compression stroke prior to ignition. Fuel injection timing can be adjusted to

control the homogeneity of the mixture. Once a homogeneous lean mixture

is achieved, the start of combustion is triggered by the high temperature

obtained near the end of the compression stroke. EGR is effective to retard

ignition timing and reduce the rate of heat release. The heat of internal EGR

can promote the evaporation of the diesel fuel. However, EGR also usually

increases HC and CO emissions. A commonly used measure of combustion

timing in HCCI research is the crank angle at which 50% of accumulated

heat release occurs. The combustion phasing crank angle at 50% heat release

is mainly adjusted through a combination of EGR rate and IVC timing with

a VVA device to reach an optimum crank angle slightly after the TDC.

The upper limit of BMEP (engine load) in HCCI operation is bounded

by the knocking combustion that is indicated by a very early ignition, a too

rapid heat release rate, an excessively high rate of cylinder pressure rise,

and loud combustion noise. In this case, a lower compression ratio and a

large amount of cooled external EGR are needed in order to control the

ignition and the combustion rate. However, the geometric compression ratio

in diesel engines is usually very high for good cold start capability. The

effective compression ratio often cannot be aggressively reduced by early or

late IVC timing in VVA because the air–fuel ratio needs to be sufciently

high. Moreover, a large amount of EGR at high BMEP will demand a very

high boost pressure and hence high peak cylinder pressure. These are the

primary reasons why HCCI has not been successfully used at high loads to

date. The lower limit of BMEP in HCCI operation is bounded by the fact that

misre occurs and the mixture fails to auto-ignite when the temperature of

the mixture of air and trapped residue gas is too low. The low-load limit of

HCCI is characterized by a large increase in CO and unburned HC emissions

as well as combustion instability. Therefore, HCCI currently has to coexist

with the conventional diffusion-controlled combustion in the same diesel

engine for different loads.

9.8.2 VVA applications for HCCI

There are two major applications of VVA for the purpose of HCCI

controls: (1) effective compression ratio control to modulate the mass

and the temperature of the charge air/gas; and (2) residue gas control to

modulate the composition and the temperature of the internal EGR. Variable

compression ratio (by IVC timing) and hot internal EGR (by exhaust valve

post-lift, intake valve pre-lift or negative valve overlap) or cooled external

EGR are used in HCCI controls of ignition timing and combustion phasing.

Regardless using internal or external EGR, intake VVA is a necessary

enabler for diesel HCCI.

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612 Diesel engine system design

VVA control for HCCI is reviewed by Tunestål and Johansson (2007).

VVA applications in diesel HCCI are reported by Babajimopoulos et al.

(2002), Milovanovic et al. (2005), Shi et al. (2005), Helmantel and Denbratt

(2006), Kodama et al. (2007), Nevin et al. (2007), Zhao (2007), Peng et al.

(2008), Murata et al. (2006, 2008), and He et al. (2008). Related fundamental

research on different VVA strategies used for HCCI has been conducted by

Caton et al. (2005a, 2005b). VVA controller design for HCCI is presented

by Strandh et al. (2005) and Bengtsson et al. (2006).

Compression ratio control with VVA for HCCI

A variable compression ratio, from the ratio as high as in the conventional

diesel engine to the ratio as low as in the gasoline engine, is desirable for

the HCCI diesel engine as the operating condition changes from cold start to

HCCI. A lower compression ratio than that in the conventional diesel engine

is usually required in HCCI in order to prevent premature or too early auto-

ignition of the diesel fuel. A true variable compression ratio (VCR) system

(Ryan et al., 2004) with sufciently fast response is ideal in HCCI controls

at different speeds and loads. However, a VCR engine is not mature enough

to be production-viable at the current time (Roberts, 2003; Cao, 2007). VVA

can be used to achieve a similar effect by reducing the effective compression

ratio at the price of decreased air–fuel ratio. (Additional information on VCR

engine designs, mainly gasoline engines can be found in Rabhi et al., 2004;

Tomita et al., 2007; Tsuchida et al., 2007; and Kobayashi et al., 2009).

VVA can change the cylinder pressure and temperature by early or late IVC

and therefore increase the usable operational range of diesel HCCI. The IVC

timing is adjusted to lower the compressed gas temperature near the TDC to

prevent too early auto-ignition and to increase the ignition delay to enhance

fuel–air mixing with prolonged premixing time to reduce smoke. However,

an excessively low effective compression ratio causes high unburned HC

emissions and fuel consumption. For example, the compression ratio cannot

be less than 5:1, as reported by Helmantel and Denbratt (2006). Note that

late IVC instead of early IVC has been successfully used in many HCCI

engines (e.g., Nevin et al., 2007; He et al., 2008).

Residue gas control with VVA for HCCI

HCCI combustion control can be achieved by adjusting the EGR rate and

temperature. The EGR gas is able to suppress the knocking combustion.

Most diesel HCCI concepts use high levels of EGR (e.g., 40–60%) and

EGR cooling along with low compression ratios. A large amount of hot

internal EGR is helpful for the low-load operation of HCCI by heating the

fresh charge and thus reducing the need for compression heating. VVA can

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613Advanced diesel valvetrain system design

achieve the controls of gas composition and EGR temperature via internal

EGR mechanisms. VVA also offers the advantage of faster response compared

to the external EGR system. The thermodynamic performance of internal

EGR by using VVA for CAI gasoline engines is reviewed by Zhao (2007)

and Furhapter (2007).

There are generally two types of VVA strategies that are used to obtain

internal EGR: (1) exhaust gas re-breathing (i.e., re-induction); and (2) residue

gas compression (i.e., trapping or retention, also called negative valve overlap

with early EVC and late IVO). The re-induction strategy refers to introducing

the exhaust gas from the exhaust port into the cylinder. The trapping strategy

refers to purposefully trapping the residue gas in the cylinder. The losses

of the re-induction strategy occur during the gas owing process across the

intake and exhaust valves and in the manifold. The trapping strategy has the

following disadvantages: thermal losses through the cylinder wall, blow-by

and extra pumping loss during the recompression. Previous experimental

research (Caton et al., 2005) revealed that the re-induction strategy gave

signicantly higher engine thermal efciency and lower NO

emissions

than the trapping strategy at the same engine speed and load, and had better

tolerance to the marginal combustion or misre at the lower load. Furhapter

(2007) reported that at low loads the trapping method had lower pumping loss

than the re-induction method, while at high loads the re-induction method

showed slightly lower pumping loss. He also concluded that for CAI gasoline

engines the internal EGR supply method with the re-induction strategy via

a secondary exhaust valve post-lift event during the late intake stroke is

the most favorable operation strategy to ensure stable auto-ignition. The

re-induction strategy with exhaust post-lift gives hotter gas than the intake

pre-lift does.

In addition to the considerations of residue gas fraction and pumping loss

during the gas exchange processes, the effect of residue gas temperature

is also important in the internal EGR strategy for HCCI controls because

it strongly affects auto-ignition. Negative valve overlap is used to initiate

auto-ignition by trapping a proportion of the exhaust gas in the cylinder.

Injecting diesel fuel in the negative valve overlap interval is another method

of forming the homogeneous mixture to achieve HCCI combustion (Shi et

al., 2005). Shi et al. (2005) pointed out that increasing the negative valve

overlap had a positive impact on the combustion stability at low loads but

impair the stability at high loads. Peng et al. (2008) discovered that a larger

negative valve overlap results in higher cylinder gas temperature, higher

temperature homogeneity, stronger turbulence intensity, and improved fuel

vaporization and air–fuel mixing to assist HCCI combustion.

Overall, although hot internal EGR is helpful for HCCI, it seems not likely

that the hot internal EGR achieved by VVA can replace cooled external EGR

in HCCI for knock intensity control (Helmantel and Denbratt, 2006).

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614 Diesel engine system design

9.9 Cylinder deactivation performance

9.9.1 Introduction of cylinder deactivation

Cylinder deactivation (abbreviated as CDA by some authors) was an old

concept originated a century ago. It is also called cylinder cutout, cylinder-

on-demand, displacement-on-demand, variable displacement, and variable

cylinder management. It refers to the technology that deactivates fuel injection

and valve operation of selected cylinders and thus exibly varies the effective

engine displacement at low loads (e.g., usually 0–5 bar BMEP).

The following can be deactivated individually or in a combined manner for

selected cylinders in different cylinder deactivation methods: spark ignition

or fuel injection, intake valve, exhaust valve, port or manifold ow. The

simplest form of cylinder deactivation is to only turn off the spark ignition

and/or fuel injection to stop the combustion process. However, the pumping

loss in this method cannot be signicantly reduced since the deactivated

cylinders still continue consuming power to pump the air through. Moreover,

the deactivated cylinders will cool quickly as the fresh air ows through

them, thus creating difculties in subsequent re-ring and emissions problems.

Although disabling either the intake or exhaust valve can prevent the engine

from pumping the air through, some pumping loss will still incur as the air

ows through the operating valves. For example, only deactivating the intake

valve causes the exhaust gas to be repeatedly inducted into and expelled out

of the deactivated cylinders.

The most effective and commonly used method of cylinder deactivation

to minimize pumping loss is to shut off both fuel injection and all the valves

(intake and exhaust) of the deactivated cylinders (Watanabe and Fukutani,

1982; Dresner and Barkan, 1989b; Falkowski et al., 2004). Sandford et al.

(1998) reported that when only the intake valve in a gasoline engine was

deactivated the fuel consumption reduction was 7%. But when both the

intake and exhaust valves were deactivated the fuel consumption reduction

was 14–17%.

The active cylinders must have higher fueling rate than they would if the

engine were run without cylinder deactivation in order to maintain the same

engine brake power. The higher load (fueling rate) means higher thermal

efciency and fewer HC emissions for the cylinder. For the conventional

gasoline engine, the higher load also means a higher intake manifold pressure and

reduced throttling and hence lower pumping loss for the ring cylinders.

The power loss in the deactivated cylinder is low, only due to a little

pumping loss of recompression, heat transfer and blow-by. In addition to

the pumping loss reduction, BSFC improvement with cylinder deactivation

comes from the following three other sources: (1) the decrease of the camshaft

friction power used to activate the intake and exhaust valves (as measured

by Watanabe and Fukutani, 1982); (2) less cylinder heat transfer loss for the

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whole engine because only some of the cylinders are ring; and (3) possible

piston ring friction power reduction due to decreased cylinder pressure in

the deactivated cylinders.

Variable engine displacement is a very attractive feature for advanced

diesel engines due to its advantages of fuel economy at part load. On the

other hand, design challenges in engine balance and NVH exist for cylinder

deactivation due to its deviated excitation forces in the cylinders compared

to those designed for the non-deactivation operation. Cylinder deactivation is

normally used in the engines that have an even number of cylinders since the

engine can operate in ring with half of its cylinders while retaining an even

ring interval. It is usually used in six or more cylinders in order to avoid

excessively high vibration. For example, in a V8 engine with the ring order

1–5–2–6–4–8–3–7, either the cylinders 1, 2, 3, and 4 or the cylinders 5, 6, 7,

and 8 can be deactivated to fulll the requirement of a constant ring interval

of 180° crank angle. The constant ring interval is achieved by deactivating

the two inner cylinders on one bank of the engine and the two outer cylinders

on the other bank. In general, from the vibration perspective, V8, V10 or V12

engines are more suitable for cylinder deactivation than V6 or I4 engines.

Cylinder deactivation has been used in limited production engines since

the 1980s (e.g., General Motor’s 1981 Cadillac Eldorado V8 gasoline engine

with two- and four-cylinder deactivation, denoted as V8–6–4; Mitsubishi’s

Orion-MD engine as presented by Fukui et al., 1983; and Mitsubishi’s two-

cylinder deactivation in an 1.6 liter I4 gasoline engine in 1992).

Almost all the published work on cylinder deactivation was conducted

on the gasoline engine. Cylinder deactivation performance was investigated

by Bates et al. (1978), Watanabe and Fukutani (1982), Dresner and Barkan

(1989a, 1989b), Hatano et al. (1993), Sandford et al. (1998), Leone and

Pozar (2001), and Douglas et al. (2005). The research conducted by Leone

and Pozar (2001) provides an excellent comprehensive summary on cylinder

deactivation for gasoline engines. The NVH of cylinder deactivation is

discussed by Falkowski et al. (2004), Lee and Rahbar (2005), and Bemman

et al. (2005). The details of design and transient controls are introduced by

Bates et al. (1978), Kreuter et al. (2001), Zheng (2001), Falkowski et al.

(2004), Stabinsky et al. (2007), and Rebbert et al. (2008). Only Watanabe

and Fukutani (1982) included a brief description of the testing results of

cylinder deactivation for a naturally aspirated diesel engine in their study.

9.9.2 Mechanisms and performance beneﬁts of cylinder

deactivation

Inuential factors of cylinder deactivation benets

The performance benets and issues of cylinder deactivation compared to

the non-deactivation operation are measured primarily by BSFC, drivability,

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vibration, noise, emissions, and exhaust temperature (for aftertreatment).

The key factors in diesel engine cylinder deactivation design include the

following:

∑ number of cylinders deactivated

∑ strategy of valve switching timing for deactivated cylinders, which is

related to the trapped gas mass in the cylinders

∑ effective area of the xed-geometry turbine used on the engine, which

is related to the trade-off between the low-load and high-load operations

in terms of the required air–fuel ratio in the cylinder deactivation

operation at part load and the required engine power capability in the

non-deactivation operation at full load

∑ VGT vane control if a VGT is used on the engine – VGT can better

t the needs of variable engine air ow rate caused by variable engine

displacement by exibly adjusting the turbine area

∑ turbine wastegate opening control capability – for example, an

electronically-controlled wastegate is able to minimize the pumping

loss and the air–fuel ratio at part load by fully opening the wastegate

in the non-deactivation operation as a low-BSFC baseline for BSFC

comparison

∑ acceptable air–fuel ratio for soot control in cylinder deactivation

∑ in-cylinder heat transfer loss

∑ piston ring friction caused by cylinder pressure

∑ valvetrain friction power consumed to drive the intake and exhaust

valves.

The factors to determine the effectiveness of cylinder deactivation for fuel

economy improvement include engine displacement, vehicle weight, driving

cycle, and the engine and VVA designs. The commonly used test cycles or

operating modes include the EPA City and Highway cycles, the New European

Driving Cycle (NEDC), and the Japanese 10–15 Modes. The BSFC benet of

cylinder deactivation reaches the greatest value for high performance vehicles

such as those using large engine displacement with six or more cylinders and

having relatively low vehicle weight and high parasitic losses. Large engine

displacement places the vehicle operating point at lower BMEP on the engine

map, and therefore more benet of BSFC with cylinder deactivation can be

obtained compared to the smaller engines. The fuel economy improvement at

low loads (e.g., 0–5 bar BMEP) is important for real world driving, especially

in LD applications. Numerous studies (e.g., Kreuter et al., 2001) show that

in LD applications the majority of the operating points in the NEDC and at

constant-speed cruising fall in the very low load region on the engine map

(i.e., 0–5 bar BMEP, at low and medium speeds). Such a light-load condition

is very suitable for cylinder deactivation.

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Valve switching strategies in cylinder deactivation

There are three methods of trapping the gas mass in the deactivated cylinders,

depending on the timing of shutting off the valves within an engine cycle: (1)

keep the minimum gas mass and almost vacuum in the cylinder; (2) induct the

cool fresh air into the cylinder; and (3) trap the hot residue gas in the cylinder.

It is desirable to stop the valve motion of the selected valves at a specic point

in the engine cycle with minimum impact on design, durability, packaging,

and cost. The shut-off timing is critical in the way that the pressure history

inside the cylinder determines the lube oil consumption and the lubrication

condition of the power cylinder components, as well as engine vibration.

Although a minimal amount of gas mass trapped produces the minimum piston

ring friction and pumping loss of the deactivated cylinders, some gas mass

and cylinder pressure need to be maintained in order to achieve an acceptable

level of lube oil consumption and lubrication condition for the cylinder by

minimizing the possibility of sucking lubricant oil via the piston ring gaps

into the combustion chamber. In addition to lube oil consumption, another

important consideration in selecting the valve switching strategy is NVH.

The engine needs to compress an appropriate amount of air mass trapped

in the deactivated cylinders to act as a damper to smooth the vibration and

speed variation of the engine.

Rapid switching of cylinder deactivation events is important for

synchronizing the cylinders in order to ensure reliable and repeatable switching

between the operating modes without uncontrolled transient conditions

that could disturb emissions or drivability. Electro-mechanical actuation

is generally faster than electro-hydraulic actuation. Fast valve deactivation

and reactivation can be realized within one engine cycle up to high engine

speeds such as 5000 rpm with electro-mechanical systems in today’s design

(Kreuter et al., 2001).

In the valve actuation sequence, either the intake or exhaust valve can

be rst deactivated. There are generally three valve switching strategies of

deactivation:

1. Deactivate the intake valve rst immediately at the end of the exhaust

stroke without relling the cylinder with fresh air and external EGR.

2. Deactivate the intake valve rst during or at the end of the intake stroke

to rell the cylinder with a certain amount of fresh air and external

EGR.

3. Deactivate the exhaust valve rst before the burned cylinder gas is

expelled out of the cylinder in order to trap the hot exhaust gas in the

cylinder.

In the rst valve switching strategy, when the piston moves down during

the intake stroke, the cylinder pressure becomes lower than the atmospheric

pressure, thus creating a high vacuum with a possible negative effect on

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lube oil consumption. Oil consumption is caused mainly by the oil loss

via the piston rings and the valve guides. The oil can be drawn up into the

combustion chamber through the joint gap areas of the top compression

ring and the second ring by the high vacuum generated in the deactivated

cylinders during the intake stroke. Note that the conventional gasoline

engine actually runs with high vacuum in the cylinders all the time at part

load under throttled operation. In the deactivated cylinders shown in some

studies, the peak cylinder pressure was usually very low (only 2–3 bar).

The cylinder pressure at the intake BDC was only around 0.2 bar absolute

pressure, and the cylinder pressure stayed below the atmospheric pressure

with vacuum during the majority of the engine cycle (Hatano et al., 1993;

Leone and Pozar, 2001). It should be noted that the piston ring pack design

may be different between the naturally aspirated gasoline engine and the

turbocharged diesel engine.

Saito et al. (1989) conrmed in their experimental work that when the intake

throttle was closed the high vacuum in the cylinder resulted in an increase

in oil consumption through the piston ring pack in a gasoline engine. They

found that the amount of lubricant oil drawn into the combustion chamber

in the engine braking condition with the intake throttle closed was about six

times as much as that at high load with the intake throttle fully open. The

oil drawn was evidenced by a large amount of oil gathered in the top land

above the top ring joint and in the second and third lands, as well as a thick

oil lm on the piston skirt. As the engine speed increased, the oil ow was

drawn up faster.

In the third valve switching strategy where the exhaust valve is deactivated

before the intake valve, the deactivated cylinder has a warmer charge due

to the trapped hot exhaust gas. The charge is gradually cooled down as heat

transfer occurs from the gas to the cylinder walls. This strategy may result

in excessively high peak cylinder pressures in the deactivated cylinders at

relatively higher engine load during the rst few seconds immediately after

the valve deactivation. The cylinder pressure quickly decays cycle by cycle

due to the heat transfer loss within several seconds to a much lower stabilized

level (Sandford et al., 1998), which is slightly higher than the pressure

level given by the second valve switching strategy. It should be noted that

recompressing the trapped gas in the deactivated cylinders produces some

pumping loss. The pumping loss of trapping hot combustion gas is higher

than that of trapping the fresh air.

A good guideline for the deactivation sequence is to nd the best balance

among the following: avoiding undesirable vacuum in the cylinder, keeping

the cylinder as warm as possible, maintaining acceptable peak cylinder

pressure for durability and engine vibration, and minimizing the pumping

loss. In many gasoline engines, the exhaust valve was often deactivated

and reactivated rst before the intake valve (e.g., Falkowski et al., 2004).

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The reason is that the gasoline engine runs at high vacuum in the cylinder

during the intake stroke at part load with intake throttle. Therefore, if a

large amount of mass (or especially hot gas mass) is desired, shutting off

the exhaust valve rst is the only solution. In the diesel engine, the situation

is different because it does not run with vacuum under throttled operation.

Therefore, the second valve switching strategy of rst shutting off the intake

valve during the intake stroke to induct cold charge can be considered.

Normally, during the extended operation of cylinder deactivation the

cylinders need to be reactivated for a short period of time in order to keep

the cylinders warm and prevent the build-up of oil deposits in the combustion

chamber (Falkowski et al., 2004). When reactivating the cylinder, the exhaust

valve usually needs to be opened before the intake valve is opened. Reactivating

the exhaust valve rst can avoid the exhaust gas from the active cylinders

being pushed back into the intake port to cause noise problems and misre

during the following combustion cycle (Kreuter et al., 2001). Reactivating

the exhaust valve rst can also avoid high recompression pressures acting on

the intake valvetrain. Moreover, the cylinders can be alternatively activated

or deactivated in order to keep the deactivated cylinders warm or prevent

them from being excessively cooled down.

BSFC improvement of cylinder deactivation

In the gasoline engine, indicated specic fuel consumption (ISFC) increases

slightly when the engine load decreases, while BSFC increases sharply due

to the large increase in pumping loss associated with intake throttling and

the larger proportion of mechanical friction (Bates et al., 1978). Even simple

fueling cut-off in the deactivated cylinders can improve fuel consumption

because it increases the fueling in the ring cylinders so that the intake

throttle opening is increased and the pumping loss is reduced. In the diesel

engine, the effect of fuel injection cut-off is negligible (Watanabe and

Fukutani, 1982).

Because a throttle valve is not necessary for load control in the diesel

engine, the effect of cylinder deactivation was believed to be less than

that in the gasoline engine by some people in the diesel engine industry.

It was believed that the benet would only come from the reduction in

valvetrain friction power and fuel injection parasitic losses. In fact, there is

a signicant benet in fuel economy with cylinder deactivation in the diesel

engine, as shown in the simulation results presented later. Watanabe and

Fukutani (1982) reported that when half of the cylinders were shut off by

deactivating the valves in a diesel engine in their experimental work, 30%

of BSFC reduction was achieved under low idle conditions. It should be

noted that if an inappropriate fuel injection compensation strategy is used at

low idle to compensate for the larger vibration amplitude and the resulting

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