Xin Q. Diesel Engine System Design

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530 Diesel engine system design

valvetrain concepts have different component weight, stiffness, conguration

height, complexity and friction loss. In order to understand the interactions

among these characteristics, it is necessary to comprehend the functions of

valve lift prole and cam lift prole, their impact on engine performance

and valvetrain dynamics, and nally the implications on design. This section

explains these topics. The research on valvetrains has a long history in the

development of the internal combustion engine. In order to understand the

advanced features of the valvetrain (including both the conventional xed

cam system and variable valve actuation or VVA systems) in modern diesel

engines, it is necessary to rst understand the fundamentals of the basic

functions and design issues related to the valvetrain. For more information

on valvetrain design fundamentals, the reader is referred to Newton and

Allen (1954), Tauschek (1958), and Wang (2007).

9.1.1 Valve lift proﬁle and valve timing

Typical engine valve lift proles of the intake and exhaust valves in the

conventional cam-driven valvetrain are shown in Fig. 9.1. The intake and

exhaust port pressures and the in-cylinder pressure are illustrated in Fig.

Desirable high-speed exhaust cam

Desirable low-speed exhaust cam

Desirable high-speed intake cam

Desirable low-speed intake cam

Maximum

exhaust

valve lift

Aggressiveness

of valve lift

Valve lift

Piston

position

More ﬂow area

on cam ﬂank

TDC

EVO IVC

IVO

EVC

Maximum

intake

valve lift

Crank angle

Valve overlap

9.1 Control points on engine valve lift proﬁles.

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531Advanced diesel valvetrain system design

Intake port pressure

Cylinder pressure

Exhaust port pressure

Pressure (bar, absolute)

Note: Zero degree is the ﬁring TDC.

–90 0 90 180 270 360 450 540 630

Crank angle (degree)

–10

9.2 Intake and exhaust port pressure pulses at full load ﬁring at 2000

rpm of an I6 engine.

There is a gas blow-down

after EVO in ﬁring hence

less air mass is trapped at

the 360° TDC

There is no gas blow-

down in motoring hence

recompression pressure

is higher

Intake ﬂow rate at

4000 rpm 30 lb/hr

fueling rate

Exhaust ﬂow rate

at 4000 rpm 30

lb/hr fueling rate

Intake ﬂow rate at

4000 rpm no-fuel

motoring

Exhaust ﬂow rate

at 4000 rpm no-

fuel motoring

IVO

EVO

IVC

Crank angle (degree)

–90 0 90 180 270 360 450 540 630 720

0.15

0.1

0.05

–0.05

9.3 Engine valve ﬂow characteristics in ﬁring and motoring.

9.2. Typical engine valve ow characteristics are shown in Fig. 9.3. Good

volumetric efciency is achieved via appropriate valve lift and valve timing

designs. It is observed that there are several control points on the valve lift

Valve ﬂow rate (kg/sec)

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532 Diesel engine system design

proles that govern the shape and the timing of the valve lift curve, and

they are (in terms of valve timing locations in crank angle): exhaust valve

opening (EVO), exhaust valve closing (EVC), intake valve opening (IVO),

and intake valve closing (IVC). The valve overlap is usually referred to by

the duration from IVO to EVC around the TDC where both the intake valve

and the exhaust valve are open. The prole control points also include the

maximum valve lift (i.e., the ‘nose lift’) for the intake and exhaust valves.

Moreover, the lift prole is controlled by the rising rate of the lift at the

opening and closing anks. The rising rate characterizes the steepness or

aggressiveness of the lift prole.

The optimum EVO timing is usually determined by an appropriate balance

in engine performance between the useable work during the expansion

stroke and the pumping loss during the exhaust stroke. Changing EVO alters

the effective engine expansion ratio and hence has a large impact on the

thermodynamic cycle efciency of the engine. An excessively advanced (or

early) EVO timing (i.e., toward the ring TDC) results in excessive loss in

indicated power and causes a penalty in BSFC. An excessively retarded (or

late) EVO timing results in high pumping loss, especially at high engine

speeds when the engine air ow rate is high, thus also causing an increase

in BSFC.

When the cylinder pressure during the blow-down process reduces to a level

that is lower than the intake manifold pressure due to rapid decompression, the

resulting ‘under-pressure’ during the exhaust stroke may produce a positive

pumping work and hence reduces pumping loss. Such a phenomenon of the

under-pressure in the cylinder at the end of the blow-down process is known

as the ‘Kadenacy’ effect (Stas, 1999; Parvate-Patil et al., 2004). This effect

may happen when the exhaust valve is opened very suddenly and rapidly.

The rapid opening results in a rarefaction wave so that a lower pressure level

may occur during the exhaust stroke to reduce the exhaust pumping loss. The

intensity of the Kadenacy effect and the cylinder pressure immediately after

EVO are affected by the rate of valve opening, the valve ow area, and the

pressure ratio across the exhaust valve. Usually, the cylinder under-pressure

after the blow-down becomes smaller as the engine speed increases. Stas

(1999) estimated that using the Kadenacy effect may provide a possible gain

in the engine thermal efciency up to 3%.

The optimum EVC timing is a balance among the following three parameters:

exhaust valve recession or piston cutout, pumping loss, and the residue gas

trapped. The valve recession or piston cutout affects the ‘K-factor’ (detailed

in Chapter 7) and combustion quality. Residue fraction is essentially hot

internal EGR. It affects the breathing quality of the engine, and reduces the

amount of fresh air and external cooled EGR that the engine can induct. A

too-early EVC timing may result in excessively high pumping loss and too

much residue gas trapped. A too-late EVC timing requires a large valve

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533Advanced diesel valvetrain system design

recession in order to avoid the valve-to-piston contact near the TDC. It may

also induce undesirable backow (reverse ow) of the burned exhaust gas

from the exhaust manifold into the cylinder. It should be noted that a later

EVC timing may reduce pumping loss due to the larger exhaust valve ow

area during the later part of the exhaust stroke.

The optimum IVO timing is determined by the balance among the following

three parameters: intake valve recession or piston cutout, pumping loss, and

the residue gas trapped, similar to the effect given by EVC. For example, a

too-early IVO timing may require a large valve recession in order to avoid

the valve-to-piston contact and may also induce backow of the exhaust gas

owing from the cylinder or the exhaust manifold into the intake manifold.

The backow reduces the breathing capacity of the engine for inducting the

fresh air. Also note that a too-early IVO timing may reduce pumping loss

due to the larger intake valve ow area during the early part of the intake

stroke. On the other hand, EVC and IVO cause different valvetrain dynamics

issues such as the recompression pressure level accumulated near the valve

overlap TDC, the intake valvetrain vibration induced by the recompression

pressure, and the exhaust valve seating velocity.

The optimum IVC timing in the conventional cam-driven xed valvetrain

is usually determined by minimizing the backow in order to maximize

volumetric efciency. Moreover, changing IVC timing alters the effective

engine compression ratio and may largely impact the thermodynamic cycle

performance and emissions of the engine. Changing IVC timing is also

the most effective way to reduce the amount of engine air ow without

incurring throttle loss as would occur using an intake throttle valve. Many

VVA mechanisms regulate IVC timing (e.g., the throttleless operation in

gasoline engines for load control, the Miller cycle or the Atkinson cycle for

lower BSFC and NO

, and the compression ratio and charge mass controls

in HCCI).

The primary design objective related to EVC, IVO, and IVC timings in

the conventional cam-driven xed valvetrain is both volumetric efciency

and BSFC, while the primary concern on EVO timing is BSFC. A higher

volumetric efciency gives a higher air–fuel ratio at the same level of intake

manifold boost pressure without the need to raise the turbine inlet pressure in

order to achieve the high air–fuel ratio. Consequently, soot can be reduced,

and this emissions benet can also be indirectly converted to a BSFC benet

through other design/calibration means.

The valve event duration (i.e., from EVO to EVC for the exhaust valve,

or from IVO to IVC for the intake valve) is also a consideration in valvetrain

dynamics and cam mechanical design, but these dynamic considerations are

usually secondary compared to the requirements of breathing performance,

especially when the engine speed is not very high (e.g., less than 4000 rpm).

Moreover, note that engine valve ow behavior changes as the engine speed

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534 Diesel engine system design

changes due to the effects of gas wave dynamics. The optimum valve timing

is a trade-off between low speeds and high speeds as illustrated in Fig. 9.1.

This is particularly true when the engine has a very wide range of operating

speeds.

The maximum valve lift and the aggressiveness of the lift rising rate are

determined by the balance among cam stress, valvetrain dynamics, and the

possible diminishing return of the effective valve  ow area at very high lift.

An example of valve and port  ow coef cient is shown in Fig. 9.4. The valve

and port  ow coef cient is highly dependent on the valve lift but nearly

independent of the pressure ratio across the valve. Note that there is a loss

in the  ow coef cient when the manifold is tested alone with the cylinder

head. The loss is due to more  ow restrictions introduced by the manifold.

The effective valve  ow area is usually calculated by the following:

VAL

,eff,e

VALV

AL,rAL

AC =AC

AC AC

AC= AC

9.1

where A

ref

can be any reference area, but usually the valve disc diameter is

used as the reference diameter. Figure 9.3 shows that the valve  ow coef cient

pro le becomes  at at high valve lift, and this means the port gradually

becomes a choking bottleneck for the  ow capacity of the cylinder head.

When this happens, any higher maximum valve lift will only provide a small

and diminishing gain on volumetric ef ciency but cause a large dif culty in

cam design and controlling the valvetrain dynamics. The interaction between

valve timing and the port  ow coef cient occurs mainly in the low lift region

0 0.05 0.1 0.15 0.2 0.25 0.3 0.35 0.4

(Intake valve lift)/(intake valve reference diameter)

0.6

0.3

Intake valve and port ﬂ ow coefﬁ cient

Improved ﬂ ow coefﬁ cient of

cylinder head measured with

intake manifold

Baseline ﬂ ow coefﬁ cient of

cylinder head measured without

intake manifold

Baseline ﬂ ow coefﬁ cient of

cylinder head measured with

intake manifold

9.4 Illustration of valve ﬂ ow coefﬁ cient.

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535Advanced diesel valvetrain system design

because the effective valve ow area at the low lift affects the backow

behavior around EVC, IVO, and IVC.

The valve lift prole also has an interaction with the turbocharger. When

a high volumetric efciency is achieved via appropriate valve timing or valve

lift prole, the same air ow rate required to meet the emissions target can be

achieved with a lower intake manifold boost pressure, and this alleviates the

burden on the turbocharger to deliver the air ow. This is especially important

for the air–fuel ratio at low engine speeds such as at peak torque. The turbine

area is basically determined based on the air and EGR ow requirements

at peak torque (if the turbine is not a exible VGT). If the valve timing is

excessively biased toward high speeds so that the volumetric efciency is

very low at peak torque, an excessively small turbine area would have to

be used to compensate for the air/EGR ow requirements. The xed small

turbine area will result in a large penalty in pumping loss at high engine

speeds.

Finally, it should be noted that valve timing and volumetric efciency

also interact with turbocharger matching via the change in engine delta P and

gas scavenging at valve overlap. Gas scavenging refers to a cooling effect

by owing the intake air from the intake manifold to the exhaust manifold

during valve overlap in order to reduce the exhaust gas temperature and the

thermal load on the cylinder head, the exhaust manifold, and the turbine.

As shown in the theoretical analysis of the engine air system in Chapter 4,

volumetric efciency is one of the key parameters that affect engine delta P

and EGR driving capability. For non-EGR engines, a negative engine delta

P can be created by selecting a large turbine area in turbocharger matching

to enable both a pumping work ‘gain’ and gas scavenging. In this case, a

large valve overlap needs to be designed accordingly in order to facilitate

the gas scavenging. However, for EGR engines the engine delta P basically

needs to be positive in order to drive EGR. In this case, gas scavenging

from the intake manifold to the exhaust manifold is impossible. Instead,

undesirable backow occurs during the valve overlap. Therefore, for EGR

engines, the valve overlap needs to be designed very small in order to ensure

high volumetric efciency. A more detailed theoretical elaboration about the

effect of valve timing on volumetric efciency and pumping loss is provided

in Section 9.7.

9.1.2 Cam lift proﬁle

The ‘cam lift’ is actually an abbreviation of the ‘follower lift at the cam side’.

The term is usually only meaningful for the pushrod valvetrain or direct-

acting overhead cam (OHC) valvetrain where the follower lift has a simple

one-dimensional translational motion driven by the cam. In some other OHC

valvetrains (e.g., with a end-pivot nger follower), the follower often has a

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536 Diesel engine system design

two-dimensional motion so that the kinematic relationship between the valve

lift and the follower lift becomes more complicated. Note that the cam lift is

not equivalent to the contour shape of the cam lobe for the case of a roller

follower. The cam contour shape refers to the radial distance beyond the

cam base circle from the cam center to any point on the cam lobe prole.

The cam lift is not equal to the radial distance because the radius of the

roller follower is not equal to zero. For the pushrod and the direct-acting

OHC valvetrains, cam lift or essentially the follower lift is usually used to

describe the intended cam prole design, and the actual contour shape of the

cam lobe can be calculated once the cam lift prole and the roller size are

specied. On the other hand, for some OHC valvetrains (e.g., with a nger

follower), it is often more convenient to directly use the actual contour shape

of the cam lobe to dene the cam prole design.

The major difference between the valve lift prole and the cam lift prole

is that the cam prole has ramps at the opening and closing ends. The ramp

height must be sufcient to accommodate the cold lash, the valvetrain elastic

compression caused by spring preload, the thermal growth of the valvetrain

under operating conditions, the dynamic deections in the camshaft caused

by the operation of adjacent valves, the wear allowance of the valvetrain

components, the normal setting tolerances, the leak-down of the hydraulic

lash adjuster (if any, for the closing ramp), and so on. The ramp height

needs to ensure the valve opens and closes on the ramp with controlled

velocities at any operating conditions to achieve low impact noise and good

durability. Using hydraulic lash adjuster may obtain precise valve timing

control at any hot or cold conditions because there is no lash. In contrast,

a mechanical lash adjuster results in variations in valve timing in cold and

hot conditions.

The aggressiveness of the cam lift prole is determined by the height and

the width of the cam’s positive acceleration peaks (humps). A higher peak

results in a narrower width of the acceleration hump and usually causes

larger vibrations of the valvetrain.

In pushrod cam design, a target valve lift prole rst needs to be

determined by the requirement of engine breathing performance. Once the

target maximum valve lift and valve event duration are determined, a cam

prole can be designed, for example with the popular ‘Polydyne’ cam design

approach. In OHC design, due to the complexity of the two-dimensional

motion of the nger follower, the cam prole cannot be directly computed

from a prescribed valve lift motion. Usually, iterations are required by using

a kinematic or dynamic model to calculate the valve lift prole based on a

designed cam contour shape, which is then compared with the desired valve

lift prole.

In engine performance analysis, usually a ‘kinematic’ or static valve lift

is used for simplicity in order to conduct a large amount of engine cycle

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537Advanced diesel valvetrain system design

simulations to optimize the required valve timing and lift pro le. The

evaluation of the more complex dynamic or vibratory valve lift pro le only

occurs at the cam design stage by using valvetrain dynamics modeling. The

de nition of the kinematic valve lift pro le is given below. If the valvetrain

elastic vibration effect is ignored, the hot non-vibratory valve lift (i.e., hot

kinematic valve lift) for a pushrod valvetrain or a direct-acting OHC valvetrain

can be calculated as follows:

VAL

hot

AVTAV

hot

hot,,hot

MR,,MR

AV,,AV

T,,T

AVTAV,,AVTAV

lN(,lN

)

lN )lN

Wl )Wl

(

Wl (Wl

)

MR,,MR

)

MR,,MR

fc– fc

N(,N

EE,,EE

CA,,CA

MR,,MR

lN(,lN

)

lN )lN

Wl )Wl

)

Wl )Wl

)

EE,,EE

)

EE,,EE

)

EE,,EE

Wl =Wl

(

Wl (Wl

(

CA,,CA

(

CA,,CA

(

CA,,CA

MR,,MR

(

MR,,MR

(

MR,,MR

)



)

–

)



(,FN(,

,FN ,

)W )

,sV,

sVTsV

gas

VAL

)

)W )

,sV,

sVTsV

(,FN(,

KKK

VAL

sVTssV

VAL

sV,,sV

Ts,,Ts

sVTssV,,sVTssV

KK +KK

9.2

where f is the crank angle, N

is the engine speed,



is the engine power,

CAM

is the cam lift as a function of crank angle, f

is the rocker arm ratio

(either simpli ed as a constant or treated as a more complicated function of

cam angle or valve lift), and c

VT,hot

is the valvetrain hot lash (for mechanical

lash adjuster only) as a function of engine speed and load. F

pre

is the valve

spring preload. F

gas,VAL

is the net gas load acting on the valve, exerted by

the cylinder gas and the port gas, as a function of crank angle, engine speed

and load. K

s,VT

is the valvetrain stiffness de ned at the valve side. K

s,VAL

the valve or valve stem stiffness. The l

CAM

(f) · f

is usually called gross

valve lift. The kinematic lift is a static net lift when the valvetrain lash (gaps

between the components) and all the elastic compressions are subtracted.

The elastic compressions of the valvetrain refer to the total amount of

compressions caused by various forces before the valve opens. It is obvious

that the de nition of valve event duration depends on how complex the

valve lift pro le is de ned. For example, some engine catalogs use 0.004

inch above the cam base circle as ‘advertised valve event duration’. Another

method often used in the aftermarket catalogs measures valve duration at

0.050 inch above the cam base circle. When specifying or comparing valve

event durations, it is important to clarify how the duration is de ned.

In equation 9.2, it is usually dif cult to estimate the valvetrain hot lash

for the mechanical lash adjusters when the temperatures of the valvetrain

components are uncertain. Moreover, the valvetrain hot lash changes at different

engine speeds and loads. In addition, the cylinder gas load also changes at

different crank angle locations. Therefore, ideally the hot kinematic valve

lift needs to be calculated at each engine speed and load. It is obvious that

such a treatment is neither convenient nor practical. For simplicity, the cold

kinematic valve lift is usually used in engine system design and performance

simulations. The cold kinematic valve lift is de ned as follows by excluding

the speed- or load-dependent factors in 9.2:

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538 Diesel engine system design

VAL

cold

AVTcAV

oldol

cold,,cold

MR,,MR

AV,,AV

Tc,,Tc

AVTcAV,,AVTcAV

()

ll()ll

()

(

ll (ll

)

MR,,MR

)

MR,,MR

fc – fc

–

CA,,CA

MR,,MR

()

ll()ll

()

ll =ll

(

ll (ll

(

CA,,CA

(

CA,,CA

(

CA,,CA

MR,,MR

(

MR,,MR

(

MR,,MR

)

9.3

where c

VT,cold

is the valvetrain cold lash set by a feeler gauge at the valve side.

The cold kinematic valve lift is regarded as a theoretical and representative

lift pro le which can be used to compare the valve timings of different

engines.

At hot operating conditions, the intake valvetrain is usually still relatively

cold, and the cylinder gas loading at IVO is very low. At full load, the exhaust

valve train is hot, and the thermal expansion of the valvetrain components

is not negligible. But the thermal expansion and the elastic compression

caused by gas loading may offset each other so that the cold kinematic valve

lift may not be too bad an assumption to represent the real lift, especially

around EVO. However, there could be large errors at EVC because the elastic

compression effect due to gas loading is much smaller there. As pointed out

by Morel and LaPointe (1994), the comparison between the simulated and

measured port pressure traces can be used to check the dynamic valve timing.

Moreover, note that usually the valvetrain elastic compression caused by the

valve spring preload is much smaller than the cold lash for mechanical lash

adjusters.

9.1.3 Valvetrain dynamics

Cam design is a highly integrated system design topic with trade-offs between

engine breathing performance and valvetrain dynamics. Valve timing, valve

event duration, and the maximum valve lift are largely determined by the

requirements of engine breathing performance. Combustion and emissions

requirements affect cam pro le design through the K-factor, valve recession

(or piston cutout), and valve overlap height (related to valve-to-piston contact).

Cam lift pro le affects valvetrain dynamics via two factors: the gas loading

acting on the valves and the inertia force excitation induced by cam acceleration.

For example, the aggressiveness of the exhaust cam closing pro le and EVC

timing affect the effective exhaust valve  ow area during the second half of

the exhaust stroke and hence impact the recompression pressure gas loading

at the valve overlap TDC that act on the intake valvetrain.

The critical design criteria in valvetrain dynamics usually include the

following: no-follow (also called separation), vibration forces acting on the

valvetrain joints (e.g., the fulcrum plate or the rocker pivot), cam stress, cam

minimum oil  lm thickness, cam  ash temperature, valve seating velocity,

and valve spring surge. The valvetrain is an elastic dynamic system with

certain stiffness and damping characteristics. Its natural frequency is affected

by the mass and the stiffness of the components. Its damping is affected

by friction and lubrication. Valvetrain vibration is excited by both cam lift

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539Advanced diesel valvetrain system design

prole and gas load. The ring gas load increases the exhaust pushrod force

at full load (Fig. 9.5), and the recompression pressure gas load increases

the vibration amplitude of the intake pushrod force at high-speed motoring

(Fig. 9.6). The pushrod vibration force is a superposition of the static valve

spring force, the dynamic forces due to valvetrain inertia and gas loading,

and the valvetrain friction force. Note that the sudden drop of the pushrod

force at the cam nose shown in Fig. 9.5 is caused by the rocker shaft friction

force. The vibratory valve lift motion may cause variations in the opening

and closing timing of the valves at different engine speeds and loads.

Valvetrain vibration may also cause no-follow between the valvetrain

components, for example at the interface between the cam and the follower,

or at the interface between the rocker arm and the valve bridge. The no-

follow is characterized by a zero force at the cam–follower contact (or called

zero cam force) during the cam deceleration phase. The cam force is the

summation of the pushrod force and the follower’s inertia force. Therefore,

the cam force usually reaches zero at a lower engine speed than the speed

at which the pushrod force reaches zero. When no-follow happens, impact

and bounce between the components occur, and the valve motion is not

well controlled. For valvetrains equipped with hydraulic lash adjusters,

no-follow may cause the lifter to pump up to ll the dynamic clearances.

If the no-follow lasts long enough, the lifter will eventually pump up to

its full extent, preventing the valve from returning to its seat and causing

a catastrophic failure of valve-to-piston contact. Valvetrain no-follow can

be controlled by a combination of appropriate cam acceleration shape, low

recompression pressure gas loading, strong spring preload and spring rate.

3.0

2.5

2.0

1.5

1.0

0.5

0.0

–0.5

–1.0

–180 –90 0 90 180 270 360 450 540

Hot lash

end

Spring

preload

Gas

load

Valve opens here

Exhaust

pushrod

force

Exhaust valve

lift (truncated

about 2 mm)

Intake

pushrod force

IVO at 320 lb

(spring preload

& rocker

friction)

Crank angle (degree)

Cylinder

pressure

Exhaust valve lift (mm)

800

700

600

500

400

300

200

100

Pushrod force (lb) or cylinder gas

pressure (psia)

9.5 Typical valvetrain pushrod force trace at low speed full load.

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