Xin Q. Diesel Engine System Design

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428 Diesel engine system design

Impact of exhaust brake gas loading on valvetrain

The potential impact of the exhaust brake operation on the valvetrain

can be summarized as follows on the gas load acting on the valvetrain

components:

1. Excessive exhaust valve oating in the intake or compression stroke.

2. Exhaust valvetrain separation near the BDC in the exhaust stroke.

3. High exhaust cam force and stress near the end of the exhaust

stroke.

4. High intake cam force at the beginning of the intake stroke.

5. Intake valvetrain separation caused by high recompression pressure.

Figure 6.7 shows that there are two exhaust pressure pulses in the early

portion of the intake stroke (360–540°) and the compression stroke (540–0°),

respectively. These pulses may overcome the cylinder pressure and the exhaust

valve spring preload so that the exhaust valve momentarily re-opens and

closes abruptly with bouncing and uncontrolled high valve seating velocity,

if the exhaust manifold pressure is too high. Moderate exhaust valve oating

is acceptable (Schmitz et al., 1992), but not excessively large. The valve

oating also causes other durability problems such as the fretting of the

valve stem, the keeper and the retainer.

Exhaust port pressure

Cylinder pressure

High recompression

pressure caused

by high exhaust

manifold pressure

These pulses of

pressure difference

can make the

exhaust valve ﬂoat

off the valve seat.

–90 0 90 180 270 360 450 540 630

Crank angle (degree)

Pressure (bar, absolute)

6.7 I6 engine exhaust port pressure pulse with conventional exhaust

brake (2600 rpm).

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429Engine brake performance in diesel engine system design

Figure 6.7 also shows that there is a high recompression pressure around

the 360∞ valve overlap TDC during exhaust braking at high engine speeds. The

high recompression pressure is caused by high exhaust manifold pressure. The

recompression pressure excites the violent vibration of the intake valvetrain

so that the intake peak pushrod force or cam force can become excessively

high just after the valve overlap TDC, and intake valvetrain separation may

occur.

Figures 6.8 shows the net gas force acting on the exhaust valve during one

engine cycle of the exhaust brake. The positive net gas force tends to push

the exhaust valve open, while the negative net gas force tends to push the

valve to close (or to increase the pushrod force). It is observed that during

the exhaust cam event, the effect of the net gas loading in exhaust braking is

to aggravate the exhaust valvetrain separation at high speeds near the BDC

(e.g., at 190∞ crank angle), and to increase the exhaust cam force and stress

at the end of the exhaust stroke. The exhaust cam stress increase in exhaust

braking is shown in Fig. 6.9. During the intake and compression strokes,

the effect of exhaust braking is to cause the exhaust valve oating, high

intake pushrod force or cam force, and intake valvetrain separation induced

by the high recompression pressure. The above valvetrain dynamics issues

limit the retarding power of the exhaust brake. Figure 6.10 summarizes these

constraints imposed by the valvetrain design.

Cam force (without gas

load) at 3000 rpm (thinner

curve) and 4500 rpm

(thicker curve)

Net gas load threshold

beyond which the exhaust

valve ﬂoats off the valve seat

Net gas load on the exhaust valve of the worst cylinder during

exhaust braking at 3000 rpm (thinner curve) and 4500 rpm

(thicker curve)

Inﬂuence of V8 exhaust-brake instantaneous pressure on

valve ﬂoating and valvetrain dynamic no-follow

0 90 180 270 360 450 540 630 720

Crank angle (degree) Note: The ﬁring TDC is at zero degree.

Force (N)

3500

3000

2500

2000

1500

1000

500

–500

–1000

–1500

6.8 Impact of V8 exhaust manifold pressure on valvetrain

performance with exhaust brake.

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430 Diesel engine system design

Exhaust valve oating analysis

A detailed analysis of the exhaust valve oating is provided as follows. The

exhaust valve oating dynamics can be analyzed based on a force balance

in the following:

Exhaust cam stress in exhaust brake

Cam stress at 3000

rpm with gas load

Cam stress at 4500

rpm with gas load

Cam stress at the

cranking speed

Cam stress at 4500

rpm without gas load

0 90 180 270 360 450 540 630 720

Crank angle (degree)

Cam stress (psi)

300,000

250,000

200,000

150,000

100,000

50,000

6.9 Impact of exhaust braking on cylinder pressure gas loading and

exhaust cam stress.

Engine retarding power at

VGT vane very closed

Valvetrain limit

(no-follow, peak force)

Exhaust valve

ﬂoat limit

Engine retarding power

at VGT vane fully open

Engine speed

Exhaust brake retarding power

6.10 Engine exhaust brake design constrained by valvetrain limits.

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431Engine brake performance in diesel engine system design

VAL

cyl



FK+ FK

lp+ lp

(–

lp(–lp

Ap(–Ap

cyl

(–

cyl

VAL

Ap(–Ap

VAL

cyl

Ap(–Ap

cyl

Sp,,Sp

VAL

cyl,,cyl

VAL

(–

lp(–lp

Ap(–Ap

cyl

(–

cyl,,cyl

(–

cyl

VAL

Ap(–Ap

VAL

Ap(–Ap

VAL

rrtr

VAL

por

VAL

por

por,,por

t,,t

VAL,,VAL

f,,f

)

AF)AF

)

AF – AF

6.12

where m

VAL

is the exhaust valve mass, l

VAL

is the lift of the  oating motion

of the valve, F

pre

is the exhaust valve spring preload, K

s,Sp

is the spring rate

(stiffness), p

cyl

is the in-cylinder pressure, A

VAL,cyl

is the exhaust valve head

area exposed at the in-cylinder side, p

port

is the exhaust port pressure, A

VAL,port

is the exhaust valve area exposed at the exhaust port side, and F

VAL,f

is the

valve stem friction force. This force balance of the exhaust valve indicates

that the valve will  oat off its seat once the net gas load overcomes the

spring preload.

As pointed out by Akiba et al. (1981) in their simulation and experimental

analysis of the dynamic motion of exhaust valve  oating in the exhaust brake

operation, the valve  oating affects engine cycle simulation results and the

predicted exhaust manifold pressure. Therefore, the simulation of the valve

 oating and bouncing motions should be interactive or coupled in nature

between the valve dynamics and the engine cycle simulation.

Exhaust runner or manifold design can reduce the peaks in the exhaust

pressure pulses that act on the exhaust valve. The reduced pulses may prevent

the large unacceptable valve  oating so that a high retarding power can be

achieved. Akiba et al. (1981) showed that exhaust manifold grouping and

the manifold cross-sectional area had a large impact on the retarding power.

There was also an effect on the pumping loss due to large cylinder-to-cylinder

variations of the exhaust pressure pulses during exhaust braking.

6.3.5 Interaction with turbocharger performance and

compression brakes

The instantaneous exhaust pressure pulses are affected by the use of

compression brakes. For turbocharged engines at low speeds where the

turbine pressure ratio is close to one, or for naturally aspirated engines,

the pulsation amplitude of the exhaust pressure pulses can be reduced if a

compression brake is used in conjunction with an exhaust brake. In this case,

the exhaust braking valve opening can be made smaller to achieve a higher

cycle-average exhaust manifold pressure by targeting for the same limit of

the instantaneous peak pulse value. As a result, a higher retarding power

can be achieved without making the valve  oating issue worse. Schmitz

et al. (1992) reported that the retarding power could be 10–15% higher by

using this technique.

The retarding performance of the conventional exhaust brake is also

complicated by turbine performance, especially at high engine speeds. When

the turbine pressure ratio is much greater than 1, if the exhaust brake is

used together with a compression brake at high engine speeds, closing the

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432 Diesel engine system design

exhaust braking valve at the turbine outlet actually results in a reduction in

turbine pressure ratio and air ow rate. As a consequence, both the engine

delta P and the retarding power decrease. When the turbine pressure ratio is

closer to 1, closing the exhaust braking valve leads to a monotonic increase

in engine delta P and retarding power. Figure 6.11 shows the simulation of

such a complex behavior of the exhaust brake and its interaction with the

compression brake.

Braking at 2600 rpm with two-stage turbo, compression-release brake turned on

Retarding power increases

Engine retarding power (hp)

Turbine pressure ratio

decreases from 4 to 1

Post-turbine

exhaust brake

valve fully open

0 1000 2000 3000 4000 5000 6000

Exhaust brake valve effective opening area (mm

)

Solid curves:

exhaust brake placed

at HP turbine inlet

Dotted curves:

exhaust brake placed

at LP turbine outlet

HP turbine

wastegate opening

13.5 mm

HP turbine

wastegate opening

6 mm

HP turbine

wastegate fully

closed

Braking at 2600 rpm with two-stage turbo, compression-

release brake turned off

Retarding power increases

Engine retarding power (hp)

Turbine pressure ratio

decreases from 2.5 to 1

Post-turbine

exhaust brake

valve fuly open

0 1000 2000 3000 4000 5000 6000

Exhaust brake valve effective opening area (mm

)

Solid curves:

exhaust brake placed

at HP turbine inlet

Dotted curves:

exhaust brake placed

at LP turbine outlet

HP turbine

wastegate

opening 13.5 mm

HP turbine

wastegate

opening 6 mm

HP turbine

wastegate fully

closed

6.11 Interaction between compression brake and conventional

exhaust brake.

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433Engine brake performance in diesel engine system design

6.4 Compression-release engine brake performance

analysis

6.4.1 Types of compression brakes

The types of compression brakes can be classied according to the following

three criteria: (1) the method of how the braking valve motion is controlled

(i.e., valve opening and closing timings precisely timed or not); (2) the

measures used to activate the valve; and (3) whether the valve actuation is

only dedicated to the retarding operation. In the conventional compression

brake, the braking valve’s opening and closing timings are precisely controlled

by mechanical, hydraulic or electromagnetic means and dedicated to the

retarding operation. In the bleeder brake, the braking valve continuously

opens by a small lift amount during either the entire engine cycle or during

the compression, expansion, and exhaust strokes without precisely timed

controls for the valve opening and closing timings. In the exhaust-pulse-

induced compression brake, the braking valve opening and closing timings

are naturally controlled by the available exhaust pressure pulses in the exhaust

port inherent in the engine (i.e., by oating the braking valve) rather than

controlled by a precisely dened mechanical motion such as in the conventional

compression brake. The exhaust-pulse brake can be further classied into

turbine-inlet pulse control brake and turbine-outlet pulse control brake. In the

VVA compression brake, the braking valve’s opening and closing timings

are controlled by a VVA device which is not dedicated only to the retarding

operation (i.e., it is also used for ring). An example of the conventional

brake is the Jake brake. Examples of the bleeder brake are Mercedes-Benz’s

decompression brake (Schmitz et al., 1992) and Navistar’s bleeder brake

used on the I6 engines. Examples of the exhaust-pulse brake are Navistar’s

(including MWM International Motores at Brazil) and MAN’s EVBec type

of brakes (e.g., Barbieri et al., 2010).

These four types of brakes have their own design characteristics and

different constraints. For example, the conventional compression brake

may have a fairly high cam force caused by the high peak cylinder pressure

near the braking TDC acting on the braking valve. This cam force and the

associated cam stress often are the limiting factor for the achievable retarding

power. On the other hand, the exhaust-pulse brake does not have this type

of loading due to its different hydraulic locking mechanism used to produce

the braking valve lift near the TDC. Therefore, it can tolerate a much higher

peak cylinder pressure thus producing a much higher retarding power. These

different types of brakes also have different noise characteristics due to the

difference in the frequency of the exhaust pressure pulses. The conventional

brake is very noisy. The bleeder brake and the exhaust-pulse brake are much

quieter (e.g., by more than 10–15 dBA).

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434 Diesel engine system design

6.4.2 Principles of compression-release brake design

Fundamental mechanisms of compression braking

The design objectives for compression-release brakes include achieving high

retarding power and low-speed torque, variable power levels, low noise,

low weight, low cost, and high reliability while satisfying all the design

constraints such as peak cylinder pressure, component loading, exhaust

manifold gas temperature, and cylinder head component metal temperatures.

Specically, the major design challenges related to braking performance

include the complexity of the following: (1) the valvetrain design for air

ow control; (2) the exhaust system design for noise control; and sometimes

(3) the mechanical loading control.

Compression brake design and performance have been sporadically reported

during the past half a century. The Jake brake design and performance were

introduced by Cummins (1966). The relationship between the turbocharger

and the compression brake was discussed by Morse and Rife (1979). Braking

valve timing/event control for improved retarding performance was addressed

by Meistrick (1992). A development process of the compression brake to

solve the design challenges at Jacobs Vehicle Equipment Company was

described by Freiburg (1994).

The design principles of compression brakes can best be understood by

an in-depth engine cycle simulation since the core of the principles is based

on thermodynamic cycles, volumetric efciency, valve ows, and gas wave

dynamics. As mentioned earlier, the retarding power of a compression brake

consists of the contributions from the indicated strokes and the pumping

loss strokes. In the indicated strokes, the concept of retarding efciency

introduced earlier clearly shows that an increase in the peak cylinder pressure

followed by a fast blow-down is highly desirable in order to increase the

retarding power. Two fundamental mechanisms should be involved in any

high-performance compression brakes. They are the compression-release

mechanism and the braking gas recirculation (BGR) mechanism (Fig. 6.12).

The compression-release mechanism has been well known since the birth

of the rst compression-release brake. The efciency of the compression-

release process at a given intake manifold boost pressure level depends

on the braking valve events (e.g., opening timing and lift prole). Figures

6.13–6.15 illustrate the compression-release process.

The braking gas recirculation (BGR) theory

BGR is a relatively new mechanism related to turbocharged engines discovered

in recent years by a group of researchers at Jacobs Vehicle Systems and the

author at Navistar (Hu et al., 1997b; Yang, 2002; Meistrick et al., 2004;

Xin, 2008). BGR is a technique to supplement the cylinder charge with the

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In exhaust-pulse-induced brake,

the braking valve lift is induced

by an exhaust pressure pulse

This is ‘braking gas recirculation’

(BGR), which inducts hot exhaust

gas into the cylinder to enhance

the ‘compression’ effect and

produce hotter gas to feed to the

turbine to generate higher intake

boost pressure

Exhaust-pulse-

induced brake (e.g.,

MANEVBec brake)

Exhaust cam

Valve lift Valve lift

Conventional

compression brake

Expansion stroke

Exhaust stroke

Intake stroke

Compression stroke

Intake cam

–90 0 90 180 270 360 450 540 630

Crank angle (degree)

–90 0 90 180 270 360 450 540 630

Crank angle (degree)

6.12 Mechanisms of compression release engine brakes.

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–180 CoMPR 0 PoWER 180 ExHAuST 360 INTAkE 540

BDC TDCF BDC TDC BDC

Crank angle (degree)

The exhaust valve does not open around TDC.

Cylinder pressure (bar)

62.5

50.0

37.5

25.0

12.5

0.0

Note: This is engine motoring without

fueling at 2100 rpm, roducing 88 hp

retarding power

Retarding

power is

produced by

‘compression’

in this stroke

In motoring, there is no

‘release’ process for the

cylinder pressure to decay

quickly. The pressure produces

positive power output in the

expansion stroke to offset

the retarding power in the

compression stroke

These two strokes produce

pumping loss power. Higher

cylinder pressure in the

exhaust stroke or lower

pressure in the intake stroke

gives higher pumping loss

Piston

moves up

Piston

moves down

Piston

moves up

Piston

moves down

6.13 Cylinder pressure trace and engine retarding power produced by pumping loss in motoring.

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–180 CoMPR 0 PoWER 180 ExHAuST 360 INTAkE 540

BDC TDCF BDC TDC BDC

Crank angle (degree)

The exhaust valve opens around TDC to ‘release’.

Cylinder pressure (bar)

75.0

62.5

50.0

37.5

25.0

12.5

0.0

Retarding

power is

produced by

‘compression’

in this stroke

Piston

moves up

Piston

moves down

Piston

moves up

Piston

moves down

Note: This is compression-release braking

at 2100 rpm, producing 408 hp retarding

power

The pumping loss in these

two strokes contributes much

less than the ‘compression-

release’ effect to the total

engine retarding power

In compression-release engine

brakes, the cylinder pressure

is released quickly to prevent

any signiﬁcant pressure left in

the expansion (power) stroke

from producing positive power

output

6.14 Cylinder pressure trace and engine retarding power produced by compression brake.

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