Xin Q. Diesel Engine System Design

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540 Diesel engine system design

These need to be designed under the constraints of the maximum cam stress

limit and the minimum penalty of engine breathing in terms of pumping loss

and volumetric efciency.

Pushrod force (lb)

Cylinder pressure (bar, absolute)

1200

1100

1000

900

800

700

600

500

400

300

200

100

1200

1100

1000

900

800

700

600

500

400

300

200

100

1600 rpm full load

1600 rpm motoring

2800 rpm full load

2800 rpm motoring

Cylinder pressure

at full load

Cylinder pressure

at motoring

Intake valvetrain pushord force test data

0 90 180 270 360 450 540 630 720

Crank angle (degree)

0 90 180 270 360 450 540 630 720

Crank angle (degree)

Intake valvetrain pushrod force test data

Cylinder pressure

at motoring

Cylinder pressure

at full load

9.6 Effect of recompression pressure gas loading on intake valvetrain

vibration.

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541Advanced diesel valvetrain system design

The performance of hydraulic lash adjusters is important for engine

breathing and valvetrain dynamics. The design and modeling information about

hydraulic lash adjusters is provided by Abell (1969), Herrin (1982), Kreuter

and Maas (1987), Phlips and Schamel (1991), Porot and Trapy (1993), Zou

and McCormick (1996), Dammers (1997), Zhao et al. (1999), and Okarmus

et al. (2008). This topic is not covered in detail in this book.

Valvetrain design is highly complex because all the parameters involved

are inter-related. They exhibit different dynamic behavior at different engine

operating conditions such as low speeds, high speeds, full-load ring, no-fuel

motoring, exhaust braking, etc. The design parameters need to be determined

simultaneously through optimization subject to the constraints of engine

performance and valvetrain dynamics. The effects of the design parameters

on valvetrain dynamics are summarized in Table 9.1.

9.1.4 Valvetrain no-follow

When a vehicle is driven downhill, it is possible for the transmission speed

to exceed the shaft speed desired for the present gear of the transmission.

Such a condition is called an over-speed condition because the transmission

over-speeds its gear. If the over-speed lasts long enough (e.g., two seconds)

the transmission gear might be damaged. Similarly, there is also a condition of

engine over-speed. Many sub-systems or components in an internal combustion

engine have a speed limit subject to mechanical failure. The valvetrain is one

of them. Above a certain engine speed, the following can happen to damage

the valvetrain or the engine: valvetrain no-follow, excessively high valve

seating velocity, high pushrod force exceeding the strength of valvetrain

joints, etc. The limiting factor for most engines is valvetrain no-follow. Severe

no-follow occurs when the pushrod force reaches zero and the duration of

the zero force lasts for more than a few degrees of crank angle.

An engine can over-speed under three conditions. First, if the engine is

provided with a sufcient amount of fuel, appropriate air, and a relatively low

load, it will accelerate to the over-speed limit on its own. Engine governors

are used to control this type of non-motoring condition by limiting fueling.

Secondly, the engine can over-speed when the vehicle is in a gear and the

vehicle load suddenly decreases to create a sufcient momentum to force

the engine speed to go above the governed speed through back-loading the

transmission. An analogy to this situation is that a marine engine over-speeds

when the ship propeller is out of the water in a rough sea. Thirdly, the engine

can be motored without fueling to the over-speed limit when an outside force

increases the engine speed. An example is that the vehicle travels downhill

with the transmission gear engaged and the operator’s foot off the accelerator

pedal. If the gravity force pulling the vehicle down the grade is higher than

the retarding force provided from the engine and the vehicle, the vehicle

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Table 9.1 General trend of valvetrain design effects on valvetrain dynamics

Recompression

pressure at high-

speed motoring

Valvetrain

no-follow

speed

Peak cam

force

Maximum

cam stress

and wear

Cam radius

of curvature

Valve

seating

velocity

Valvetrain

friction

Valvetrain

noise

Valve

spring

surge

Reduce valve (or

cam) event duration

–

(worse)

–

(worse)

Advance EVC timing +

(worse)

–

(intake

worse)

(intake

worse)

(intake

worse)

Retard IVO timing +

(intake

better)

–

(intake

better)

–

(intake

better)

Reduce valve overlap

size

(worse)

–

(intake

worse)

(intake

worse)

(intake

worse)

Increase exhaust

manifold pressure

(worse)

–

(intake

worse)

(intake

worse)

(intake

worse)

Increase maximum

valve or cam lift

–

(worse)

–

(worse)

Reduce rocker arm

ratio

–

(worse)

–

(better)

–

(better)

–

(worse)

Reduce valvetrain

weight

(better)

–

(better)

–

(better)

–

(better)

–

(better)

–

(better)

Increase valvetrain

stiffness

(better)

–

(better)

–

(better)

–

(better)

–

(better)

–

(better)

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Increase the

aggressiveness

of valve lift slope

with higher cam

acceleration

–

(better)

–

(worse)

(exhaust

worse at full

load)

–

(worse)

Reduce valve spring

preload or spring rate

–

(worse)

–

(better)

–

(better)

–

(better)

Reduce cam base

circle diameter

–

(worse)

–

(worse)

Reduce roller

follower’s roller

diameter

(worse)

–

(worse)

Note: ‘+’ sign indicates increase. ‘–’ sign indicates reduction. ‘Intake’ indicates intake valvetrain only. ‘Exhaust’ indicates exhaust valvetrain

only.

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544 Diesel engine system design

speed and the engine speed will increase. Over-speed can occur especially

when the driver downshifts to a too low gear (e.g., shift from fourth gear

at 45 mph vehicle speed and 3000 rpm engine speed to third gear at 4500

rpm) or makes a skipped downshift in a manual transmission (e.g., shift

from fourth gear to  rst gear directly).

Some vehicles and automatic transmissions are equipped with over-speed

warning or prevention systems to prevent engine over-speed by disallowing

downshifting the transmission to certain gears according to the driving

condition. In addition, some systems with manual transmissions can alert

the driver to the danger of motoring the engine beyond the over-speed limit

during gear selection so that the operator is offered an opportunity to take

corrective actions to avoid damaging the engine due to excessive motoring.

The engine needs to be designed to possess a good mechanical capability

up to the over-speed limit.

A target valvetrain no-follow speed is established in engine design as one of

the over-speed limits of the engine. It is often used as the maximum allowable

operating speed of the engine. An appropriate no-follow speed target is a

compromise between the requirement of engine breathing capability and the

requirement of protecting engine over-speed in manual transmission vehicles.

The no-follow speed target should be selected above the high-idle speed with

a certain margin. An excessively low no-follow speed target would produce

the valvetrain separation problem within the normal operating speed range

of the engine. On the other hand, an excessively high (or over-designed) no-

follow speed target would force either the cam acceleration hump to be too

low or the valve spring force to be too high. A low cam acceleration results

in a low valve  ow area and high pumping loss. A high spring load leads to

high friction force and poor BSFC. Moreover, it should be noted that although

the selection of the valvetrain no-follow speed target is based largely on

the normal motoring condition, the more severe but less frequent operating

conditions such as engine braking or transients should also be considered.

9.1.5 Valvetrain inertia force

Valvetrain inertia force is affected by component mass and valve acceleration.

In the simpli ed single-degree-freedom valvetrain dynamics model (detailed

later in Fig. 9.15), the inertia force at the valve side can be calculated as

the product of the valvetrain equivalent mass and valve acceleration. The

valvetrain equivalent mass usually can be calculated as follows:

valv

etaine

dge

mmªmm

mm mm

valv

mm+ mm

AAA

RARARAR

ValveSid

pus

9.4

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545Advanced diesel valvetrain system design

where m

valve

is the valve mass, m

retainer

is the retainer mass, m

bridge

is the

valve bridge mass, m

spring

is the valve spring mass, m

pushrod

is the pushrod

mass, m

is the rocker arm mass, l

RA,ValveSide

is the rocker arm length at the

valve side, f

is the rocker arm ratio, and l

RA,0

is the radius of gyration. The

effective mass of the rocker arm can be considered as the mass of the rocker

arm concentrated at one point which is at a distance l

RA,0

from the center

of rotation, i.e.,

Im= Im

where I

is the moment of inertia of the

rocker arm with respect to the center of rotation. Lower cam acceleration,

higher stiffness, lower weight, and higher natural frequency of the valvetrain

result in lower vibration amplitude of the valve acceleration. The design

guideline for valvetrain inertia force control is that the design needs to

achieve a precise target of valvetrain no-follow speed with an optimized gas

load. Neither over-design nor under-design is acceptable for a good balance

between breathing performance and valvetrain dynamics.

9.1.6 Valvetrain gas loading and recompression

pressure

The gas loading acting on the intake and exhaust valves has a great impact

on valvetrain dynamics and cannot be ignored. The gas loading acting on

the valve face is the in-cylinder pressure. The gas loading acting on the

valve back is the pressure in the port. The mechanisms and effects of gas

loading are different for different valves and at different engine operating

conditions. The gas loading is affected by the valve  ow characteristics of

the engine (Fig. 9.3). In general, there are three issues in valvetrain dynamics

related to gas loading: (1) the peak cam (or pushrod) force for both intake

and exhaust valvetrains; (2) intake valvetrain no-follow; and (3) exhaust

valve  oating.

In engine  ring and especially at full load, the exhaust valve needs to

overcome a large in-cylinder pressure before it opens. The gas loading at EVO

at full load has a dominant impact on the exhaust cam force. The cam stress

at EVO at full load may sometimes become even higher than the cam stress

at the cam nose during low-speed cranking. Once the exhaust valve opens,

the cylinder pressure gradually decays with a strong blow-down process to

expel the gas mass out of the cylinder. As a result, there is only very little

gas mass trapped at the end of the exhaust stroke so that the recompression

pressure at the valve overlap TDC is usually very low at full load. Therefore,

the net gas loading acting on the intake valvetrain at full load is usually very

low.

On the back of the exhaust valve, there are usually several exhaust port

pressure pulses (in the case of pulse turbocharging) within an engine cycle

to try to push the valve to open. The exhaust valve spring preload and the

in-cylinder pressure counteract the exhaust pressure pulses. At full load,

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546 Diesel engine system design

the intake manifold pressure is high. Therefore, the pressure differential

between the exhaust port pressure pulse and the in-cylinder pressure during

the intake stroke is usually not large enough to oat the exhaust valve off

the valve seat. Figures 9.5 and 9.6 illustrate the effects of gas loading on

pushrod vibration.

In engine motoring, the in-cylinder pressure at EVO is low and the net gas

loading acting on the exhaust valvetrain is very low. However, due to the lack

of blow-down, a large amount of in-cylinder residue gas can be trapped in the

cylinder and compressed to a very high recompression pressure at the valve

overlap TDC (Fig. 9.6). A high recompression pressure is caused directly by

a high exhaust manifold pressure, which results from the following: (1) high

engine speed; (2) small turbine area; (3) small effective exhaust valve ow

area at the cam closing side due to a small valve, a low port ow coefcient,

an early EVC timing, or a less aggressive exhaust cam lift prole; and (4)

exhaust braking. Note that an early EVC timing, may result in a small valve

overlap, which usually needs to be used in modern high EGR engines. The

recompression pressure has a sharp decaying characteristic just past the TDC

when the piston moves away from the TDC. The high pressure level together

with its inherent ‘shock loading’ characteristics make the recompression

pressure the most dominant factor for intake valvetrain vibration in modern

diesel engines. The gas loading and cam acceleration induce violent intake

valvetrain vibration at high-speed motoring. The vibration can be so severe

that the rst peak of the cam (or pushrod) force can be even higher than

the maximum cam force of the exhaust valvetrain. The peak force is often

followed by a severe no-follow condition (i.e., a long duration of zero pushrod

force as shown in Fig. 9.7). Due to the recompression pressure, valvetrain

no-follow usually occurs rst in the intake valvetrain rather than the exhaust

valvetrain.

In exhaust braking, the exhaust manifold pressure is raised to a very high

level in order to increase pumping loss and retarding power. The intake

manifold pressure usually remains low. The recompression pressure in exhaust

braking can be much higher than that in motoring. Severe intake valvetrain

no-follow may occur if not designed properly. Moreover, the large pressure

differential across the exhaust valve between the exhaust port pressure pulse

and the in-cylinder pressure during the intake stroke may oat the exhaust

valve off the valve seat.

During transient acceleration, the EGR valve is usually closed in order to

ow more fresh air into the cylinder to match the suddenly increased fueling

amount. The increased turbine ow rate due to EGR valve shut-off causes

an immediate large increase in the exhaust manifold pressure. However, the

intake manifold pressure cannot respond instantly to quickly build up due

to turbocharger lag. As a result, there is a large engine delta P which may

create two problems for the valvetrain: (1) intake no-follow due to high

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547Advanced diesel valvetrain system design

recompression pressure; and (2) exhaust valve oating during the intake

stroke.

The problem of excessively high peak exhaust cam force can be solved

using a smaller rocker arm ratio or designing a stronger structural strength.

The problem with exhaust valve oating can be solved or alleviated, if not

totally avoided during transients or engine braking, through a proper selection

of valve spring preload. Recompression pressure and intake valvetrain no-

follow controls are the most sophisticated issues in valvetrain system design

since they require a system-level approach to optimize the solutions in the

early stage of the engine design. There are several methods to reduce the

recompression pressure at high-speed motoring by either reducing the exhaust

manifold pressure or bleeding off the in-cylinder gas mass: (1) choosing

a large turbine area; (2) opening the VGT vane or the turbine wastegate;

(3) using large exhaust valves; (4) using less restrictive exhaust ports; (5)

retarding the EVC timing (i.e., using larger valve overlap in cam design);

or (6) using a more aggressive exhaust cam closing lift prole (i.e., larger

valve ow area at the closing side). Moreover, there are several methods to

sustain a high recompression pressure in order to control the intake valvetrain

no-follow: (1) using a less aggressive intake cam lift prole (i.e., lower cam

acceleration); (2) using a longer intake cam duration by retarding IVC; (3)

retarding IVO to reduce intake vibration; or (4) using a stronger intake valve

spring at the expense of higher cam stress and valvetrain friction.

Turbocharger selection is largely determined by the boost requirement in

the ring operation. Modern low-emissions EGR engines tend to use two-

Cylinder pressure (psia) and pushrod load (lb)

1400

1300

1200

1100

1000

900

800

700

600

500

400

300

200

100

–100

270 300 330 360 390 420 450 480 510 540 570 600 630 660 690 720

Crank angle (degree)

Comparison between acceptable no-follow (A) and unacceptable

no-follow (B), 4400 rpm no-fuel motoring

Recompression pressure A

Vibration of a good cam

under recompression

pressure trace A

recompression pressure B

Vibration of a bad cam under

recompression pressure trace B

9.7 Valvetrain no-follow and its control with reduced recompression

pressure.

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548 Diesel engine system design

stage turbocharging with a very small turbine at the high-pressure stage.

Such a conguration may generate higher exhaust manifold pressure and

recompression pressure than a single-stage turbocharger. Consequently,

the intake valvetrain no-follow control becomes a more challenging task.

Typically, opening the VGT vane at high-speed motoring (e.g., 4000 rpm)

may reduce the recompression pressure by approximately 25 psi (1.72

bar).

Using late EVC timing and large valve overlap has been the classical

method to reduce the recompression pressure and pumping loss in the past

for non-EGR engines. Typically, every 1° crank angle retardation of the

exhaust cam closing lift prole leads to a 5 psi (0.34 bar) reduction in the

recompression pressure. The disadvantages of this method in EGR engines

include the higher residual fraction trapped in the cylinder (especially at low

engine speeds), lower volumetric efciency, and the larger valve recession

required in order to avoid valve-to-piston contact. A lager valve recession

results in a larger dead volume in the combustion chamber and higher soot

emissions. More ‘squared’ exhaust cam closing prole combined with an

early EVC timing can be considered as an alternative solution. If carefully

designed, this method may avoid the penalties associated with the simple

EVC retardation. However, the aggressiveness of the cam closing lift

prole and the level of recompression pressure reduction are limited by

the minimum allowable negative radius of curvature of the cam and the

maximum allowable exhaust valve seating velocity. Compared to a normal

exhaust cam, the specially designed aggressive exhaust cam may reduce the

recompression pressure by 10–20 psi (0.69–1.38 bar) at an increased risk of

high valve seating velocity.

The intake valve opening timing has a negligible inuence on the

recompression pressure. However, the intake valve opening timing affects

the intake valvetrain vibration signicantly due to the change in the location

where the peak recompression pressure acts on the cam acceleration curve. A

well-optimized smooth acceleration prole of the intake cam may sustain 30

psi (2.1 bar) recompression pressure at the same intake no-follow condition

without a signicant degradation in engine breathing performance. However,

there is a slight penalty of higher cam stress at the cam nose. Typically,

every 2° crank angle increase in the intake cam duration (e.g., by retarding

IVC timing) results in the capability of sustaining an additional 12 psi

(0.83 bar) recompression pressure at the same no-follow condition. The

disadvantages with this IVC-retarding approach include negative impact

on cold start capability and the degradation of volumetric efciency at low

engine speeds.

In summary, recompression pressure and intake valvetrain no-follow

controls require an integrated solution and a balanced trade-off among cylinder

head design, turbocharger selection, combustion chamber design, exhaust

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549Advanced diesel valvetrain system design

cam design, and intake cam design. The appropriate design sequence is as

follows:

1. Reasonably establish a target no-follow speed based on engine application

needs (e.g., in engine–transmission matching) without over-design.

2. Optimize the cylinder head design to achieve the maximum exhaust

valve size and exhaust port ow coefcient.

3. Optimize turbocharger conguration, turbine area and wastegate valve

opening and their control strategies at high engine speeds.

4. Determine the allowable exhaust valve recession and the corresponding

dead volume in the combustion chamber.

5. Design a good exhaust cam prole to maximize the closing-side ow

area.

6. Design a good intake cam prole to achieve smooth cam acceleration.

7. If all the above measures cannot solve the no-follow problem, the intake

valve spring preload or spring rate has to be increased with increased

cam stress or reduced maximum intake valve lift.

Among these design measures, the exhaust and intake cam prole designs

have a vital importance. Figure 9.7 illustrates an example of intake valvetrain

no-follow control by reducing the recompression pressure and designing a

superior and smooth acceleration prole of the intake cam.

9.1.7 Valvetrain design criteria

The valvetrain basic competitive benchmarking parameters may include the

following: valvetrain natural frequency; aggressiveness of cam acceleration;

cam base circle radius; valve spring load; valve overlap duration; effective

valve ow area; and the ratio of valve diameter to bore diameter.

The most important design criteria for the valvetrain in turbocharged EGR

diesel engines are summarized as follows:

∑ low valvetrain weight and high stiffness (i.e., high natural frequency)

∑ coordinated design of cam prole and turbocharging to achieve high

volumetric efciency in a wide speed range with appropriate maximum

valve lift, large effective valve ow area, and proper valve timing

∑ small valve overlap (i.e., short overlap duration and low overlap lift height)

for low internal residue fraction and good transient acceleration

∑ no valve-to-piston contact but designed for tight clearance at the TDC

∑ low recompression pressure at the valve overlap TDC achieved by

appropriate turbine area, exhaust cam lift prole, and EVC timing

∑ smooth shapes of cam acceleration and jerk, appropriate cam ramp

height

∑ coordinated design of cam lift prole, valvetrain dynamics and engine

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