Xin Q. Diesel Engine System Design

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438 Diesel engine system design

hot gas recirculated from the exhaust manifold during engine retarding in

order to provide higher retarding power. It is essentially an ‘internal EGR’

technology used in the retarding operation. As shown earlier in Fig. 6.7 in

the discussions on the exhaust brake, the engine inherently generates several

exhaust pressure pulses in the exhaust port. These pulses try to open the

exhaust valve during the intake or compression stroke to ow the exhaust

gas into the cylinder. BGR utilizes this phenomenon. In BGR, the braking

valve is opened in late intake stroke until the early part of the compression

stroke to induct the hot exhaust gas into the cylinder via a pressure differential

between the exhaust port pressure pulse and the in-cylinder pressure. The

braking valve lift event for BGR can be generated by any of the following: a

mechanical cam, a bleeder event, a VVA device, or by the exhaust-pressure-

pulse-induced free motion of the braking valve such as in the case of the

exhaust-pulse-induced brake. The BGR valve lift event is sometimes also

called a ‘secondary lift’ in braking. BGR can help increase the turbocharger

speed and achieve high peak cylinder pressure. In the open literature, the

BGR concept has not been fully explored, and its mechanism has not been

clearly explained. The BGR theory is developed as follows.

High peak cylinder pressure is the key for high retarding power. The

following factors can increase the cylinder pressure: high engine compression

ratio, high intake manifold boost pressure, and a large amount of gas mass

trapped in the cylinder during the compression stroke. The boost pressure

is related to turbocharger operation. The trapped gas mass is related to the

–90 0 90 180 270 360 450 540 630

Crank angle (degrees)

Compression-release engine

brake at different engine speeds

Cylinder pressure (bar, absolute)

200

150

100

Firing

Positive power

Negative

power

Exhaust

brake

valve

Intake

valve

0 180 360 540 720

Crank angle

Valve lift

Cylinder pressure (bar, absolute)

200

150

100

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1

Instantaneous cylinder volume/maximum cylinder volume

6.15 The engine cycle process of compression-release brake.

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439Engine brake performance in diesel engine system design

valve events used in each stroke and the associated reverse ows (if any)

as well as the engine volumetric efciency. The intake boost pressure can

be increased by many design measures (e.g., using higher turbocharger

efciency or a smaller turbine area such as a VGT, or reducing the turbine

outlet pressure). One important and effective method of boosting the ow is

to increase the turbine inlet temperature. Higher exhaust energy or exhaust

manifold gas temperature received by the turbine can increase the turbine

speed to deliver high intake boost pressure. The key issue is where to obtain

this high exhaust temperature for the turbine. BGR conducts an ‘exhaust gas

recovery’ to harvest the hot exhaust gas mass back into the cylinder, and

compresses the hot charge by the piston to even hotter at the TDC, and release

it to the turbine. Note that the hot exhaust mass is essentially originated from

the moving vehicle’s kinetic or potential energy during the braking process

when the engine brake works like an air compressor. Transferring this energy

from the vehicle to the turbine inlet to speed up the turbine is inherently a

very efcient mechanism. With the blow-down process this hot exhaust gas

is fed into the turbine as high pressure and high temperature pulse energy

ow to increase the turbocharger speed. As a consequence, the boost pressure

becomes higher, and the cylinder charge is compressed to an even higher

temperature. Such a compounding effect continues until the entire charging

process reaches an equilibrium condition (e.g., steady state). The key enabler

for this highly effective cylinder-pressure building process is BGR. A higher

BGR valve lift event with appropriate BGR valve timing will increase the

retarding power signicantly. Analysis and experimental work have shown

that BGR can increase the retarding power of the compression brakes to

extremely high (i.e., much higher than the ring rated power), provided the

high cylinder pressure does not create structural design or stress problems on

the braking components (depending on the type of the compression brakes).

In summary, any efcient compression brake design should be directed toward

the two important fundamental braking mechanisms: the compression-release

process, and the BGR process.

For four-stroke engines, the right BGR valve lift location is the crank

angle durations in the late intake stroke and the early compression stroke

where the intake valve is almost closed and the exhaust port pressure is

higher than the in-cylinder pressure (i.e., around 500–600° after the ring

TDC). For different engines (I4, I6, with divided or undivided turbine entry or

exhaust manifold) and at different speeds, the exhaust port pressure pulsation

pattern can be different. The optimum BGR valve lift location can change

accordingly.

Engine air ow rate affects the in-cylinder gas temperature and the

exhaust manifold gas temperature. In general, the lower the air ow rate, the

higher the temperatures. A large air ow rate is desirable in engine braking

because it cools the cylinder head components, the injector nozzle tip and

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440 Diesel engine system design

the exhaust manifold for better durability. It should be noted that increasing

the turbine outlet pressure causes a reduction in the turbine power and the

air ow rate.

Also note that because the air ow rate is high in BGR operation, the

increased exhaust manifold gas temperature is usually still below the

allowable temperature limit. In fact, it can be lower than the high exhaust

temperatures encountered in the exhaust brake operation where the air ow

is choked. The air ow rate and the boost pressure are related via the engine

volumetric efciency. It should be noted the engine volumetric efciency

in compression braking with BGR becomes complex due to the change in

the valve events, compared to the case in ring operation. Considering the

requirements on both retarding power and component thermal durability,

a good design objective for compression brake design can be stated as to

achieve simultaneous high engine air ow rate and high exhaust manifold

temperature within the design constraints. BGR is the mechanism to

achieve this goal in a balanced manner. The ultimate retarding power limit

is bounded by allowable peak cylinder pressure and exhaust manifold gas

temperature.

Turbocharger matching for engine braking

Turbocharged diesel engines have different retarding power capability

compared to the naturally aspirated diesel engines owing to their different

levels of peak cylinder pressure and engine delta P. The engine delta P is

governed mainly by turbine area and engine air ow rate. The volumetric

efciency is related to valve events. It should be noted that a xed-geometry

turbine or a wastegated turbine selected for good ring operation is usually

inferior to a VGT for good retarding power capability. Therefore, both ring

and retarding operations need to be considered in engine air system design

and turbocharger selection/matching.

Figure 6.16 illustrates the engine breathing characteristics and turbocharger

matching. It shows that if the engine volumetric efciency can be altered

by valve events, the turbocharging performance in engine braking can be

adjusted for a specic brake design purpose. It also shows an illustration

of the methods of improving engine retarding power relative to the critical

parameter – the engine delta P. The engine delta P is not only related to the

pumping loss and retarding power, but also to the design constraints such

as exhaust valve oating, the exhaust valve spring preload, and the exhaust

manifold pressure. Different design strategies, either low or high engine delta

P, can be selected in engine brake designs. Figure 6.17 summarizes the root

causes and the corresponding design solutions for the common problems

related to compression brake performance.

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441Engine brake performance in diesel engine system design

Summary of key design principles for compression brakes

The key elements in a powerful compression brake are summarized as

below:

∑ Achieve high intake manifold boost pressure and peak cylinder pressure.

This ensures a powerful compression-release process and a high air

ow.

∑ Do not throttle turbine outlet at high engine speeds where the turbine

pressure ratio should be much greater than 1. This ensures a high turbine

speed with a large pressure ratio to deliver high intake boost pressure.

∑ Use BGR. The retarding power of the conventional compression brake

is usually limited by the peak cylinder pressure that can be sustained

by the maximum allowable braking cam stress, while the exhaust-pulse

compression brake does not have such a cam stress limit because its

Engine volumetric efﬁciency

Engine retarding power

Compressor pressure ratio

Naturally aspirated engine, intake

manifold pressure is approximately

equal to ambient pressure

Engine breathing

characteristic line at

different speeds

Rely on strong

compression-release effect

to increase retarding power

Compressor

efﬁciency

contours

Turbocharged engine, intake

manifold pressure is higher

than ambient pressure

Engine air ﬂow rate Engine air ﬂow rate

Engine delta P

(exhaust manifold pressure minus intake manifold pressure)

The slope is roughly

the reciprocal of

volumetric

efﬁciency

Rely on high engine delta P and

pumping loss to increase retarding

power

Higher volumetric

efﬁciency by

valve timing

Lower

volumetric

efﬁciency

6.16 The principle of compression-release engine brake.

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Low

retarding

power

Low

engine

delta P

Low boost

pressure, low

turbine power

Excessively

high

injector tip

temperature

Low pumping loss

Low air ﬂow rate

Root cause

High potential

Low potential

High potential

Replaced by

modulating

turbine area or

entry or a valve

at turbine inlet

Ineffective brake

valve lift proﬁle

Better valve actuation

for higher brake lift

Reduce EGR valve or

wastegate leakage

use divided turbine

entry

Raise design limit

of engine delta

P by 1–2 bar and

reduce turbine

area by VGT or

block one entry

in twin-entry or

install a valve at

the turbine inlet

Design better

hydraulics/mechanism

EGR valve or turbine

wastegate leakage

undivided turbine

entry

Engine delta P

design limit too low

Large turbine area

Low turbine inlet

pressure

High turbine outlet

pressure

Low turbo efﬁciency

(e.g., compressor choke)

Design solution Effect on

retarding power

Injector temperature

design limit too low

Low compression-

release effect

Turbocharger lag

Reduce turbo lag

High heat losses in

exhaust manifold

Delete or fully

open turbine outlet

exhaust ﬂap valve

Change turbo

aerodynamic design

Increase turbine

inlet temperature

6.17 Logic ﬂowchart of engine brake performance problems and design solutions.

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443Engine brake performance in diesel engine system design

braking valve lift is generated by the exhaust pressure pulses. Moreover,

the exhaust-pulse brake naturally achieves the BGR effect. The maximum

lift of the BGR event in the conventional brake is limited by the peak

cylinder pressure. The exhaust-pulse brake does not have this limit and

is more advanced. The key to make the exhaust-pulse brake successful

is to produce all the braking valve lift events by throttling the turbine

inlet rather than the turbine outlet. Because both the intake boost

pressure and engine delta P can be high in the exhaust-pulse brake, the

resulting required high exhaust manifold pressure should be managed

carefully.

∑ Generate a braking valve lift prole around the braking TDC and a

BGR valve lift with a cost-effective design without the excessive design

constraints of braking cam stress due to the peak cylinder pressure.

The above theory illustrates that, when turbocharger matching and valvetrain

design are conducted, the requirements from engine braking performance also

need to be considered. It is emphasized again that the engine performance

design must be integrated as a whole system, not only because of the

complex interactions between different subsystems, but also for the purpose

of coordinating the needs of different applications.

6.4.3 Performance characteristics of compression

brakes

The retarding power of compression brakes is sensitive to many design and

operating factors including engine speed, total engine displacement in braking

(or the number of cylinders), engine compression ratio, braking valve effective

ow area, braking valve lift/timing/duration, and turbine area.

The optimum braking valve effective ow area for the maximum retarding

power may vary with the engine speed. Usually, a high engine speed and

a large air ow rate demand a large valve ow area. Using one valve in

braking can reduce the gas load acting on the braking components (e.g.,

the injector, the exhaust pushrod and cam). In order to provide a sufcient

braking valve ow area, one-valve braking needs to have basically twice the

valve lift used in two-valve braking, thus demanding a larger valve-to-piston

clearance. One-valve braking and two-valve braking designs are elaborated

by Price and Meistrick (1983).

The optimum braking valve opening timing also depends on the engine

speed. The optimum timing occurs earlier at higher speeds in order to assure

adequate blow-down of the cylinder charge. Figure 6.18 presents the simulation

results by using GT-POWER for a conventional compression brake with

one-valve braking for an I6 heavy-duty diesel engine. The optimum closing

timing (i.e., re-set timing) of the braking valve after the TDC following the

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Valve lift (mm)

Cylinder pressure (bar, absolute)

150

100

120

110

100

–90 0 90 180 270 360 450 540 630

Crank angle (degree)

–90 0 90 180 270 360 450 540 630

Crank angle (degree)

BVo = –40 deg

BVo = –20 deg

BVo = 0 deg

BVo = 20 deg

Engine retarding power (hp)

Engine retarding power (hp) or

engine air ﬂow rate (lb/min)

400

300

200

100

–60 –50 –40 –30 –20 –10 0 10 20 30 40

BVo timing (degree crank angle)

–60 –50 –40 –30 –20 –10 0 10 20 30 40

BVo timing (degree crank angle)

VGT vane

opening

= 0.5

VGT vane

opening

= 0.65

VGT vane

opening

= 0.75

VGT vane

opening

= 1.0

VGT vane

opening

= 0.5

VGT vane

opening

= 0.65

VGT vane

opening

= 0.75

VGT vane

opening

= 1.0

Retarding power

Air ﬂow rate

6.18 Conventional compression brake performance sensitivity at 2100 rpm with one-valve braking.

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445Engine brake performance in diesel engine system design

blow-down may yield an additional gain in retarding power because closing

the valve as the piston moves down in the expansion stroke may minimize the

cylinder pressure (see Fig. 6.19 for the illustration of the instantaneous valve

ows in a bleeder brake simulation). Figure 6.20 presents a comprehensive

simulation analysis on the effect of bleeder valve lift, turbine-out ap valve

opening, VGT area, and engine speed for a bleeder brake. This analysis

also reveals the interaction between the compression brake and the exhaust

brake.

Superior low-speed retarding torque has been a design challenge. From a

vehicle drivability point of view, running at 1500 rpm instead of 2100 rpm

during braking is probably easier for the driver because down-shifting is

not needed. Low-speed braking can also reduce compression brake noise.

Using VGT, BGR or a turbine-inlet exhaust brake may help increase the

low-speed torque. At very low engine speeds where the turbine pressure

ratio is low or close to 1, a turbine-outlet exhaust brake can be considered.

Note again that the throttling effects at the turbine inlet and outlet with an

exhaust brake are fundamentally different. Throttling the turbine inlet does

not affect the turbine pressure ratio or the turbine speed as drastically as

throttling the turbine outlet where the effect of increased exhaust restriction

is produced. A careful design balance between the brake, the engine, and

the turbocharger is required in order to fulll all the design requirements.

The retarding performance of different compression brakes combined with a

variable exhaust brake is discussed by Hu et al. (1997b). More information

Air velocity at bleeder valve (m/s), lift 0.02 inch

Air velocity at bleeder valve (m/s), lift 0.03 inch

Air velocity at bleeder valve (m/s), lift 0.04 inch

Air velocity at bleeder valve (m/s), lift 0.05 inch

0 180 360 540 720

Exhaust bleeder

braking valve lift

Intake valve lift

Crank angle (degree)

Exhaust ﬂow velocity (m/s)

800

600

400

200

–200

–400

–600

160

140

120

100

Valve lift (mm)

6.19 Illustration of braking exhaust valve ﬂow velocity in bleeder

brake at 2100 rpm.

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dkfgope

1.5

20 40 60 80

Exhaust ﬂap opening (degree)

20 40 60 80

Exhaust ﬂap opening (degree)

20 40 60 80

Exhaust ﬂap opening (degree)

20 40 60 80

Exhaust ﬂap opening (degree)

Bleeder braking performance with VGT area = 570 mm

Bleeder braking performance with VGT area = 700 mm

Retarding power (hp) 2800 rpm

Retarding power (hp) 1800 rpm

Exhaust bleeder valve lift in bleeder

braking (mm)

Exhaust bleeder valve lift in bleeder

braking (mm)

Exhaust bleeder valve lift in bleeder

braking (mm)

Exhaust bleeder valve lift in bleeder

braking (mm)

1.5

0.5

1.5

0.5

1.5

0.5

1.5

0.5

220

200

180

160

140

300

250

200

150

6.20 Engine bleeder brake

performance sensitivity to

the effects of bleeder lift,

turbine-outlet exhaust ﬂap

valve, engine speed, and

VGT opening area.

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447Engine brake performance in diesel engine system design

on the design effects on braking performance is provided by Schmitz et al.

(1992) and Imai et al. (1996).

A detailed comparison between engine motoring, the conventional

compression brake and the bleeder brake on their instantaneous working

processes within an engine cycle is illustrated in Fig. 6.21. The corresponding

steady-state performance data are shown in Table 6.1. This example shows

that engine cycle simulation plays a very powerful role in understanding the

principles of the engine brakes.

6.4.4 Design constraints of compression brakes

The key durability design constraints of the compression brake are summarized

as follows. These design limits need to be examined carefully.

∑ Peak cylinder gas pressure, which affects engine structural durability.

∑ In-cylinder component metal temperature, for example the injector nozzle

tip temperature. The temperature of the injector tip during engine braking

is very different from that during the ring operation where a supply

of fuel ow to the injector provides a cooling effect. The injector tip

temperature depends directly on the in-cylinder heat ux, which is affected

by the engine air ow rate and the compressor boost pressure.

∑ Exhaust manifold gas temperature, which is affected by BGR and engine

air ow rate, which is in turn inuenced by the turbocharger performance

and the exhaust braking valve opening at the turbine outlet (if any).

∑ Valvetrain gas loading and component stress. These include the gas

loading from the peak cylinder pressure near the braking TDC acting on

the braking valve, the pushrod, and the cam. The braking load and the

stress on the cam and the rocker arm may exceed the valvetrain loading

capability in the ring operation. The loading on the valve system in

the compression brake is discussed by Imai et al. (1996).

∑ Engine delta P. This limit is evaluated to prevent excessive exhaust valve

oating. It is related to the exhaust spring preload and the associated

exhaust cam stress.

∑ Braking house stress and structural/hydraulic compliance.

∑ Coolant heat rejection. Although the heat rejection of the compression

brakes is far less than that of a hydrodynamic retarder, the coolant

heat rejection may become signicant at very high retarding power. It

is necessary to check the heat rejection to ensure a sufcient cooling

capacity.

∑ All other common design constraints that should be considered in the

engine system design for ring operation, for example turbocharger

speed, compressor outlet air temperature, etc.

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