Xin Q. Diesel Engine System Design

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590 Diesel engine system design

Larsen (1991), Asmus (1991), Kreuter et al. (1992), Nuccio and Marzano

(1992, a review), Gallo (1992), Ke and Pucher (1996), Anderson et al.

(1998), Pischinger et al. (2000), Pierik and Burkhard (2000), and Ribeiro

and Martins (2007).

Gasoline engine valve overlap

In the gasoline engine, valve timing has traditionally been designed to optimize

the operation at the high-speed full-load conditions with the intake throttle

wide open. The valve overlap size in the gasoline engine is determined based

on the considerations of high-speed power, low-speed torque, residue gas

quantity at part load for charge composition control and emissions, charge

loss to the exhaust at the full load, and idle quality. At part load, if the engine

is throttled a large pressure difference between the exhaust manifold and

the intake manifold (e.g., high vacuum) is created. Then, a large amount of

residue gas (internal EGR) is inducted from the exhaust manifold into the

cylinder through the valve overlap.

A small valve overlap in the gasoline engine leads to the following: (1)

low volumetric efciency at high engine speeds due to poor scavenging;

(2) high torque at low speeds; (3) low internal EGR at part load; (4) low

backow of unburned mixture of fuel and air to the exhaust system; and (5)

good idle quality due to low residue gas fraction. The low backow to the

exhaust duct avoids high BSFC and also protects the catalyst. The improved

idle stability and quality may enable a reduction in the idle speed to reduce

idle fuel consumption.

Gasoline engine IVC and EVO

Either early or late IVC can reduce the effective compression ratio of the

engine and the compression pressure. The load control without throttling in

the gasoline engine is achieved by using an early or late IVC (or variable

valve lift) to reduce the mass of charge trapped in the cylinder and maintain

the intake manifold pressure near the atmospheric pressure rather than a

high vacuum. Therefore, the pumping loss can be greatly reduced. Note

that Stone and Kwan (1989) pointed out that at low loads the reduced

compression ratio due to IVC change may reduce the benet of throttle loss

reduction. Early IVC can be achieved by advancing IVC timing to reduce

the intake valve event duration and sometimes accompanied by a reduction

in the maximum valve lift to control the valve acceleration for the reduced

valve event duration.

Retarding EVO timing may increase the effective expansion ratio. The EVO

timing can be especially retarded to around the BDC at low speeds/loads since

the pumping loss during the exhaust stroke is not adversely affected at the low

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591Advanced diesel valvetrain system design

gas ow rate. Moreover, if the opening rate of the exhaust valve lift is higher

than the cam-driven lift, EVO can be retarded to reduce BSFC even at high

speeds by increasing the expansion work without a penalty of pumping loss

during the exhaust stroke. From the above discussions on the valve timing,

it is obvious that VVA is very important for the gasoline engine.

Performance benets of gasoline engine VVA

The performance benets of gasoline engine VVA can be summarized as

follows:

∑ Improved torque curve. The gasoline engine usually has a wide speed

range. A high volumetric efciency provides a high air ow rate and

hence more fuel can be burned to increase the power. VVA enables

optimized valve timing and high volumetric efciency at all speeds to

create higher and atter torque curve from low to high speeds. VVA is

especially useful for non-turbocharged engines.

∑ Reduced part-load throttle losses and fuel consumption. The benet of

pumping loss reduction with VVA to replace the intake throttle is most

signicant at the low-speed light-load conditions where the throttling is

the heaviest. VVA can also reduce the effective compression ratio without

a corresponding change in the expansion ratio so that the thermodynamic

cycle efciency can be improved. A range of 15–20% BSFC reduction

by VVA on gasoline engines has been widely reported and conrmed

(e.g., Dresner and Barkan, 1989b; Pischinger et al., 2000).

∑ Improved idle stability and fuel consumption. Low-idle operation largely

affects the gasoline engine fuel economy. The highest percentage of

fuel economy improvement by VVA usually occurs at low idle, and

decreases as the engine speed or load increases. The reduced small valve

overlap offered by VVA improves the idle stability and enables an idle

speed reduction to reduce fuel consumption by around 30%. Ma (1988)

reported that a 200 rpm reduction in idle speed from 800 rpm translated

to a 6.1% improvement in the ECE-15 drive cycle fuel economy. Early

IVC further reduces the pumping loss at low idle due to throttleless

operation.

∑ Emissions reduction through the controls on both cylinder charge mass

and composition (residue gas). VVA largely reduces NO

by increasing

residue gas fraction or internal EGR, for example via large valve overlap.

A small valve overlap may reduce HC and CO emissions. VVA can

also reduce HC and CO emissions during the cold start and warm-up

transients by using late IVO. Moreover, late EVO may reduce emissions

and increase the exhaust gas temperature.

∑ Emissions reduction through charge motion control. The objective of

charge motion control is to improve mixture formation and increase

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592 Diesel engine system design

internal cylinder ow for better homogeneity of the stoichiometric

mixture to accelerate combustion. Variable swirl ratio and high intake

valve ow velocity at low speeds can be achieved by low or uneven

valve lifts, valve deactivation, or negative valve overlap (i.e., EVC being

earlier than IVO around the TDC).

9.7.6 Diesel engine VVA performance

Overview of diesel engine VVA

The history of diesel engine VVA can be traced back to the 1960s when

the Jacobs Manufacturing Company used the technology of precision valve

actuation to actuate the compression-release brake. The hydro-mechanically

created braking exhaust valve event in the compression brake can be seen

as an early form of VVA technology for the diesel engine. An example of

diesel VVA on the intake valve in the ring operation was Caterpillar’s

ACERT system created several years ago that used hydraulic actuators to

change the intake valve motion and timing to achieve charge control and

internal EGR. Since recently, intake VVA has been extensively explored

in diesel HCCI combustion. VVA affects effective compression ratio and

mixture reactivity with residue gas control and temperature control to enable

HCCI combustion.

The need to exibly regulate the engine valve ows in the diesel engine

was recognized by previous researchers decades ago (Mardell and Cross,

1988). Previous research on diesel VVA performance can be categorized in

the following areas:

∑ The effects of valve overlap size and reverse ows at high loads and

low loads and for transient load response were simulated by Charlton

et al. (1990, 1991) for a non-EGR turbocharged diesel engine that had

negative engine delta P.

∑ Hot internal EGR by VVA was researched by Desantes et al. (1995),

Benajes et al. (1996), Schwoerer et al. (2004), Parvate-Patil et al. (2004),

and Millo et al. (2007).

∑ Valve timing and engine breathing performance were analyzed by

Ghaffarpour and Baranescu (1995), Lanceeld et al. (2000), Lanceeld

(2003), Parvate-Patil et al. (2004), and Deng and Stobart (2009). Lanceeld

et al. (2000) and Lanceeld (2003) studied cooled external EGR used

together with VVA on EVO and IVC timing controls with xed IVO

and EVC timings.

∑ The diesel Miller cycle with adjustable IVC timing was researched by

Ishizuki et al. (1985), Bolton and Assanis (1994), Stebler et al. (1996),

Edwards et al. (1998), Mavinahally et al. (1996), Kamo et al. (1998),

Wang et al. (2005), Millo et al. (2005), Ribeiro and Martins (2007), and

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593Advanced diesel valvetrain system design

Imperato et al. (2009). Other research on the Miller cycle (mainly for

gasoline engines) was conducted by Hitomi et al. (1995), Okamoto et al.

(1997), Clarke and Smith (1997), Ge et al. (2005), Ribeiro and Martins

(2005), Ribeiro et al. (2006), Al-Sarkhi et al. (2002, 2006), Chen et al.

(2007, 2008), Junior (2009), and Wik and Hallbäck (2008), for large

marine diesel engines).

∑ The rotary valve used for diesel VVA to achieve early IVC timing was

investigated by Ishizuki et al. (1985).

∑ VVA and turbocharger interaction was studied by Ke and Pucher (1996,

for a turbocharged gasoline engine), Edwards et al. (1998), Tai et al.

(2002), Lanceeld (2003), and Yang and Keller (2008). Gu (1995) and

Yang et al. (2009, 2010) reported using intake VVA (the Miller cycle)

combined with variable exhaust and fuel injection timing on non-EGR

diesel engines. Their system was originally called the Gu-system named

after Professor Hong-Zhong Gu’s name (Gu, 1995).

∑ General performance benets of VVA on diesel engines were investigated

by Mardell and Cross (1988), Fessler and Genova (2004), Leet et al.

(2004), and Gehrke et al. (2008).

∑ The effects of VVA (intake valve lift and IVC timing) on combustion,

emissions, and turbulence swirl were computationally investigated by

Stephenson and Rutland (1995) and Munnannur et al. (2005).

∑ The applications of VVA on diesel engine brake were discussed by Hu et

al. (1997a, 1997b), Israel (1998), Schwoerer et al. (2002), Yang (2002),

and Fessler and Genova (2004).

∑ Diesel VVA design and electro-hydraulic simulation were conducted by

Gehrke et al. (2008), Bernard et al. (2009), Hass and Rauch (2010).

∑ Intake VVA not only provides benets on diesel engine air system

performance and pumping loss reduction, but also improves combustion

efciency and reduces emissions through higher charge density, colder

combustion temperature and better in-cylinder charge motion. The

research on the combustion benets of diesel intake VVA was conducted

by Kim and Kim (2002), Su et al. (2005, 2009a, 2009b), Kim et al.

(2009), Murata et al. (2010), Su (2010), and De Ojeda et al. (2010).

Valve overlap, IVC, and EVO in diesel VVA

The cam phasing VVA devices are widely used in gasoline engines to change

the valve overlap. Automotive high-speed diesel engines have a very high

geometric compression ratio that prevents the use of large valve overlap

offered by the cam phasing devices due to the concern of valve-to-piston

contact under the tight clearance at the TDC. Any valve pocket in the piston

top or valve recession in the cylinder head that may prevent valve-to-piston

contact is generally associated with a sensitive inuence on combustion and

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594 Diesel engine system design

emissions. At part load or low idle, the engine delta P in the diesel engine

is not as great as in the conventional gasoline engine due to throttleless

operation. Note that the gasoline engine has a much higher exhaust manifold

pressure than the intake manifold pressure (vacuum) at low idle. The

diesel engine is operated with a much higher air–fuel ratio than that in the

gasoline engine so that it can tolerate high EGR at low idle. The effect of

trapped residue mass on diesel combustion is not as detrimental as that in

the spark-ignition combustion which is burned stoichiometric. Therefore, a

large valve overlap in the diesel engine does not cause the severe problems

of high residue fraction and combustion instability that are encountered in

the gasoline engine.

The valve overlap in the non-EGR turbocharged diesel engine is a

compromise between the low-load reverse residue ow and the full-load

component temperatures. The non-EGR diesel engine is usually designed to

have large valve overlap in order to enhance gas scavenging with negative

engine delta P. Good scavenging and cooling can be achieved to reduce

the thermal load on components such as the exhaust valve, the combustion

chamber, and the turbine. At lower speeds or loads, the boost pressure is

likely lower than the exhaust manifold pressure so that a small valve overlap

is desirable in order to prevent or reduce the exhaust reverse ow. The small

valve overlap was especially important for old engines that burned heavy

diesel fuels. The exhaust residue reverse ow in those old engines was very

contaminating for the intake port and manifold. Although the use of VVA

for valve overlap control is not as critical as in the gasoline engine, VVA

can be helpful for the non-EGR diesel engine in order to achieve short valve

overlap at low loads and long overlap at high loads by properly matching the

overlap with engine delta P at different loads or speeds. A comprehensive

simulation on the effect of valve overlap is given in Fig. 9.12(a).

On the other hand, the modern EGR diesel engine is usually designed

to have small valve overlap at any speed and load in order to prevent the

residue gas backow and maximize volumetric efciency. The valve overlap

needs to be small due to the need for a positive engine delta P to drive the

external EGR ow. Note that intake manifold pressure, engine delta P and

BSFC are affected by valve overlap size (Fig. 9.12(a)). Two special cases

of valve overlap in VVA when both the intake and exhaust valves remain

open are the intake secondary pre-lift event during the exhaust stroke and

the exhaust secondary post-lift event during the intake stroke. They are

used to induce internal EGR (residue gas). The secondary valve lift events

may offer some opportunities for the diesel engine to regulate air–fuel

ratio and engine delta P. Moreover, it should be noted that a negative valve

overlap (i.e., EVC earlier than IVO) is generally not desirable due to its

negative impact on the reduced valve ow area, high pumping loss, and

high recompression pressure. The negative valve overlap can be considered

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595Advanced diesel valvetrain system design

in HCCI combustion control by using hot internal EGR (detailed in Section

9.8).

The IVC timing in diesel VVA is usually related to the so-called Miller

cycle (Miller, 1947; Miller and Lieberherr, 1957) with the turbocharging

effect. A comprehensive simulation on the effect of IVC timing is given in

Fig. 9.12(b) and (c). The EVO timing in diesel VVA is also closely related

to the characteristics of pulse turbocharging. These topics are detailed in

later sections.

The valve lift rate is limited by cam design, mechanical stresses and other

dynamic issues. In general, ow losses across the valve can be reduced by a

faster valve opening or closing rate. Camless VVA may offer an advantage

over the cam-driven VVA with a faster lift rate. However, it should be noted

that excessively fast lift rate for intake valve opening/closing and exhaust

valve closing is not necessary due to the fact that near the beginning of

the intake stroke and the end of the exhaust stroke the piston velocity and

thus the valve ow velocity are very low, and an excessively fast valve lift

(as designed in some camless devices) only provides negligible benet on

pumping loss reduction. Therefore, there is no need to pursue a near ‘square

shape’ valve lift prole for the intake valve in the camless design, especially

at low engine speeds. A trapezoidal prole is usually sufcient because the

shapes of the instantaneous piston velocity and the valve ow rate as a

function of crank angle are close to trapezoidal.

On the other hand, unlike EVC, IVO, and IVC, a fast opening lift at

EVO is usually desirable in order to achieve retarded EVO timing for better

BSFC. Moreover, faster opening of the exhaust valve lift also reduces the

valve ow losses during the blow-down process and increases the energy

available at the turbine. As pointed out by Stone and Kwan (1989), rapid

exhaust valve opening is especially needed at low engine speeds where the

turbine power is low.

Performance benets of diesel engine VVA

The traditional belief was that the diesel engine may benet less from VVA

in engine breathing performance compared with the gasoline engine for the

following reasons. First, the load control in the diesel engine is achieved by

adjusting the fuel quantity instead of throttling the air. Therefore, the potential

of using VVA to reduce pumping loss at part load is much less. Secondly,

the fuel injection equipment and the heterogeneous diesel combustion restrict

the rated speed of the diesel engine usually below 4500–5000 rpm, which is

much lower than the rated speed of the gasoline engine. Therefore, the speed

range is much smaller and the trade-off in valve timing is not so severe.

However, it should be noted that the expectation on the percentage fuel

economy improvement should be different for gasoline and diesel engines

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596 Diesel engine system design

because the diesel engine is largely used in heavy-duty applications where

even a small percentage of fuel economy gain is valuable for a large eet.

Therefore, any small BSFC gain offered by VVA should not be simply

discarded in diesel engine design. Moreover, most importantly, it is believed

that the Miller cycle with VVA can provide signicant BSFC improvement

for the modern diesel engines that have high engine delta P and high EGR.

Beyond that, there are many other benets that VVA can offer to the diesel

engine. A cost–benet assessment and a comparison with the competing

technologies are required to justify the commercial use of VVA for production

diesel engines. The competing technologies for fuel economy improvement

include hybrid powertrain, waste heat recovery, and advanced turbocharging.

The benets of diesel engine intake and exhaust VVA are summarized as

follows.

Increased low-speed torque

Increased low-speed torque can be achieved by the following: (1) retarding

EVO timing past the BDC with a penalty in BSFC (Tai et al., 2002); (2) using

the Curtil exhaust VVA system (Curtil, 1998) to utilize the exhaust reverse

ow at the compression BDC (Israel, 1998; Fessler and Genova, 2004); and

(3) advancing IVC or EVO timing (Lanceeld et al., 2000). The approach

of retarding EVO can greatly increase the air–fuel ratio due to an increase

in turbine power so that more fuel can be burned to obtain higher torque at

low speeds. The approach works well only at low speeds (e.g., 1000 rpm).

The turbine power increase is due to a higher pressure pulse at the turbine

inlet that is caused by both pulse turbocharging and a higher residue gas

pressure inside the cylinder. The EVO retarding needs to be conducted with

strong pulse turbocharging, and the compressor may surge under vey low

ow rate and high boost pressure. Therefore, careful turbocharger matching

is required in order to realize the high torque. Moreover, note that the very

late EVO gives greater expansion work but more pumping loss and increased

residue gas trapped. The associated BSFC penalty can be small in terms of

the contribution to the overall driving-cycle fuel economy due to very little

time spent at low speed full load (e.g., 1000 rpm) during acceleration.

Reduced pumping loss and fuel consumption

Reduced pumping loss is achieved primarily through the following three

measures: (1) use the Miller cycle with early or late IVC to reduce both

volumetric efciency and engine delta P while increasing boost pressure (large

benet); (2) use camless VVA to achieve fast valve lift rate to maximize

the effective valve ow area (small benet); and (3) deactivate both fuel

injection and valve ows in cylinder deactivation (large benet at low loads).

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597Advanced diesel valvetrain system design

Improved fuel consumption can also be achieved by the following means

related to the engine cycle processes: (1) optimize the variable EVO timing

(small benet); and (2) obtain better thermodynamic cycle efciency by using

a very high geometric compression ratio, a relatively larger expansion ratio

and a smaller effective compression ratio such as in the Miller or Atkinson

cycles (large benet). Previous research indicated that an improvement of

at least a few in BSFC is possible by optimizing the EVO and IVC timings

(without using the Miller cycle or cylinder deactivation yet) as well as using

fast valve lift rate with VVA, for example, 4% at low idle, 1–3% at light

load, 1% at full load, 3–5% in driving cycles. It is found that city driving

offers the largest gain in BSFC reduction percentage, and highway driving

probably offers the least gain in BSFC reduction.

Reduced emissions

Increased air ow or air–fuel ratio can be achieved to reduce soot by optimizing

the volumetric efciency (mainly through adjusting IVC) across the entire

engine speed–load domain. The gain in volumetric efciency can be several

percentage points. Effective compression ratio can be reduced by early or

late IVC to obtain lower compression pressure, combustion temperature,

and NO

. Early IVC timing also provides certain limited benet of internal

in-cylinder charge cooling due to the expansion toward the intake BDC.

Hot internal EGR (residue gas), although not as effective as cooled external

EGR for NO

reduction, can reduce NO

in certain applications that are not

subject to stringent emissions regulations. Yang and Keller (2008) reported

that late IVC could reduce NO

by 24% and BSFC by 1% at 2000–3000 rpm

and 4–5 bar BMEP for a LD small diesel engine. The charge composition

control via internal EGR is achieved by an exhaust valve post-lift during the

intake stroke (the preferred approach) or an intake valve pre-lift during the

exhaust stroke (less desirable). Leet et al. (2004) estimated NO

reduction

of 10% or 0.13 g/(hp.hr) by using VVA to achieve the 1.18 g/(hp.hr) NO

level for HD diesel engines. Millo et al. (2007) achieved 13% reduction

in NO

for a small off-road diesel engine by using internal EGR without

remarkable detrimental effects on fuel consumption and soot emission. As

reported by Fessler and Genova (2004), the Euro-IV emissions limit of

can only be reached with cooled EGR and is not achievable by using

VVA alone. As to the particulate matter (PM) and unburned HC emissions,

late EVO timing results in longer in-cylinder burning time for PM and HC

so that they can be reduced, although the pumping loss may increase with

the retarded EVO timing. Late IVO may enhance the air ow velocity and

the turbulence to reduce unburned HC emissions. In summary, VVA itself

is not a sufcient enabling technology to meet the most stringent NO

and

PM emissions regulations.

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598 Diesel engine system design

Air motion control and swirl/tumble modulation

Increased in-cylinder gas motion can be achieved by using low or unequal

intake valve lifts to increase the air ow velocity across the valve seat in

order to enhance the swirl formation in the cylinder and reduce PM and

BSFC at low speeds. Another alternative is to open the intake valve late

after the TDC to create a large pressure differential across the intake valve,

which results in high valve ow velocity and also high pumping loss. The

VVA can replace the use of swirl aps and may have comparable inuence

on the in-cylinder charge motion as the cylinder head design offers. The

VVA valve deactivation system may be more effective and obtain a higher

swirl ratio than the port deactivation system can do.

Eliminating air control valves

It is possible to eliminate some air control valves by using VVA to achieve

a similar or better function of air ow and boost pressure controls. The air

control valves include intake throttle, exhaust throttle, turbine wastegate,

and EGR valve. Eliminating the redundant air control valves can greatly

simplify the engine system design and reduce cost. VVA can regulate the

engine air ow and the pressures by using the valves in the cylinder head

with crank angle resolution without incurring the additional throttle losses

at other air control valves.

Enabled advanced combustion

Intake VVA is an enabler for advanced combustion concepts such as HCCI

(detailed in Section 9.8). VVA provides controls on variable compression

ratio and charge composition and temperature (via internal EGR) that are

required for HCCI ignition and combustion. VVA also provides opportunities

for controlling the combustion processes of the engine that use multiple

types of fuels.

Improved cold start

Better cold start means quick start, reduced HC emissions and white smoke,

and the possibility of eliminating the glow plug to reduce cost. Cold-start

quality is affected by the air temperature in the cylinder at the end of the

compression stroke, geometric compression ratio, and IVC timing. Quick

cold start can be achieved by advancing IVC toward the BDC to avoid the

reverse air ow and increase the effective compression ratio, the trapped

gas mass, and the compression temperature at the TDC. Engine geometric

compression ratio is largely determined by the cold-start requirement, and

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599Advanced diesel valvetrain system design

the compression ratio is usually higher than the ratio that would be optimally

required for the best BSFC during hot engine operations after cold start.

Unlike the low- or medium-speed diesel engines used in the applications

of marine, railway transportation, and power generation sets, automotive

high-speed diesel engines use high compression ratios. VVA with early

IVC timing also enables a lower geometric compression ratio for better

peak cylinder pressure control at rated power. VVA can also tolerate lower

quality fuels (e.g., cetane number or rating less than 40) or colder ambient air

temperatures in terms of cold-start capability. Other means of assisting quick

cold start include the following: (1) using cylinder deactivation to increase

the cranking speed; (2) opening the engine valve during the compression

stroke to motor the engine with minimum compression at the very early

stage of cranking in order to reach the cranking speed more quickly with

less energy provided by the starter motor; and (3) increasing the charge

temperature by inducting the warm gas from the exhaust port by using a

post-lift of the exhaust valve. After cold start in a cold climate, VVA can

reduce HC emissions during warm-up by using internal EGR and cylinder

deactivation. Because unburned HC emissions are low in the diesel engine,

any improvement in HC emissions via VVA is limited to cold start.

Fast warm-up

The cold start portion of the HD transient test cycle becomes more important

as the emissions regulations become more stringent. The NO

aftertreatment

system also requires fast warm-up after cold start in order to achieve high

conversion efciency. The following measures can speed up the engine

warm-up: (1) throttling the intake or exhaust valve ow by using reduced

valve lifts to add load (pumping loss) to the engine; and (2) using hot internal

EGR (residue gas) to heat the intake charge. Early EVO timing can help

speed up the warm-up of the aftertreatment system.

Reduced turbocharged lag and improved transient response

High low-end torque and fast transient response can improve drivability. Both

early EVO and very late EVO can increase turbocharger speed and reduce

turbocharger lag during acceleration transients although BSFC usually becomes

worse due to the loss of expansion work or increased pumping loss.

Improved aftertreatment performance

Intake or exhaust throttling may reduce air–fuel ratio and increase the

exhaust temperature for the aftertreatment system (e.g., DPF regeneration,

SCR, LNT). Exhaust throttling generally produces less BSFC penalty than

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