Xin Q. Diesel Engine System Design

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408 Diesel engine system design

retarding products including the Jacobs Engine Brake, Jacobs Exhaust Brake

and Jacobs Driveline Brakes. It should be noted that there are many different

types of compression brakes such as the conventional compression brake,

the bleeder brake, and the exhaust-pulse-induced compression brake. Both

JVS and other companies produce many different types of engine brakes

and driveline retarders. For example, in addition to the Jacobs Engine Brake,

several compression brakes were made by other companies. These include

Mack’s Dynatard (Greathouse et al., 1971), Cummins’ C-brake, Mercedes-

Benz’s ‘Konstantdrossel’ brake, and Pacbrake’s Engine Brake.

6.1.5 Braking mechanisms of retarders

The hydrodynamic retarder converts the mechanical energy of the drive

shaft to the heat of the uid inside the retarder via a rotor-stator design. Its

retarding torque depends on the amount of uid ow and the pressure of

the viscous uid inside the retarder. The retarding torque is controlled by

adjusting the hydraulic pressure. The uid under viscous damping is cooled

by the engine coolant through the vehicle radiator or a separate cooler (in

the case of a retarder-equipped trailer).

The electromagnetic retarder uses a disc rotating in a magnetic eld to

generate the eddy current which produces the braking torque. The braking

toque is controlled by adjusting the exciting current.

Caterpillar’s retarder ‘BrakeSaver’ (Darragh, 1974) used engine oil as

a working medium churning between a rotor and a stator. It was mounted

between the engine crankshaft and the ywheel in order to utilize the

speed–torque multiplication by the transmission.

The retarding power of an engine brake consists of the contributions

from mechanical rubbing friction power, engine accessory power (such as

the fuel pump, the water pump and the oil pump), the pumping loss power

during the intake and exhaust strokes, and the indicated power during the

compression and expansion strokes.

The simplest form of engine braking is to use the engine’s motoring

power (without fueling). The motoring power is basically a second order

function of the engine speed. Downshifting the transmission gear (e.g., from

5th gear to 4th) or turning off overdrive (in an automatic transmission) may

increase the engine speed thus increase the retarding power. However, the

engine speed should not exceed its maximum allowable limit. The engine

mechanical friction here does not include the parasitic losses from certain

vehicle accessories such as the air compressor, the power steering pump,

the radiator fan, etc. The retarding power from these vehicle accessories is

a part of the vehicle’s natural retarding capability (see equation 6.1).

In the intake-throttle brake (also called ‘vacuum engine brake’), high

pumping loss or retarding power is achieved by the fact that the piston has

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409Engine brake performance in diesel engine system design

to resist the high vacuum created in the cylinder when the piston moves

down during the intake stroke. One example of the intake-throttle brake

is that the gasoline engine develops retarding power by closing the intake

throttle plate which restricts the amount of intake air. Gasoline engines

commonly have an intake throttle to regulate the air ow to match the fuel

ow in order to achieve stoichiometric combustion. Note that the non-EGR

old diesel engines did not have an intake throttle. When the intake throttle

is closed during engine braking, a gasoline engine may generate higher

pumping loss and higher retarding power than a diesel engine of the same

displacement without an intake throttle or an exhaust brake, if compared at

the same engine speed. It should be noted that gasoline engines are designed

to run at much higher engine speeds than diesel engines and hence they may

develop higher retarding power due to their higher speed.

Similar to the intake-throttle brake, in the conventional exhaust brake

with an exhaust throttle valve, high pumping loss and retarding power

are achieved by the fact that the piston has to overcome the high exhaust

manifold pressure during braking when the piston moves up in the exhaust

stroke. The maximum potential of the intake-throttle brake comes from only

1 bar maximum in engine delta P (i.e., the difference between the 1 bar

absolute pressure in the exhaust manifold and zero bar vacuum in the intake

manifold), while the potential of the exhaust brake is much higher than 1

bar in terms of engine delta P. This is the reason why the exhaust brake is

usually used instead of an intake-throttle brake in diesel engines. In fact,

the exhaust brake is the most widely used retarder to date. It is usually used

with mid-range-power diesel engines in a wide range of Class-2 through

Class-7 vehicles, for example from the light pickup trucks that haul heavy

loads (e.g., 10,000–15,000 lbs GCVW) in hilly terrain to heavy trucks and

buses.

A compression-release engine brake is a device that can convert a diesel

engine from a power-producing machine to a power-absorbing machine, for

example to a reciprocating (piston) air compressor. During engine braking,

with the fuel ow terminated, the power required to compress the air in the

cylinder comes from the kinetic or potential energy of the moving vehicle and

the turning engine crankshaft. In this case, the engine valves remain closed

during the compression stroke. If the exhaust valve also remains closed during

the expansion stroke (as in the normal valve event in ring operation), the

compressed air will return the positive power to the vehicle via the crankshaft

during the expansion stroke so that the earlier absorbed (negative) power

is cancelled out. In this situation, the only retarding power offered by the

engine is the pumping loss during the intake and exhaust strokes as well as

the mechanical friction from the moving parts. If the exhaust valve is opened

near the end of the compression stroke or in the very early portion of the

expansion stroke, the compressed air is allowed to discharge to the exhaust

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410 Diesel engine system design

manifold thus the energy stored in the compressed air is no longer retained

in the cylinder during the expansion stroke to produce positive power. As

a result, the net effect is that the engine becomes a power-absorbing air

compressor to use the kinetic or potential energy of the moving vehicle

to pump the ambient air through the engine cylinders and then discharge

the hotter air to the atmosphere. The negative power produced during the

compression and expansion strokes contributes greatly to the total retarding

power of the engine. In this operation, such an exhaust valve is called an

engine braking valve.

In fact, the braking valve can be the exhaust valve (like in the Jake

Brake) or an additional decompression valve in the cylinder head (like in

Mercedes-Benz’s Konstantdrossel engine brake and Mitsubishi’s Powertard

engine brake). There are numerous designs and inventions to realize the

optimum functions of the braking valve(s) through mechanical, hydraulic,

and electromagnetic valvetrain mechanisms.

Although some different types of retarders can be combined to use on one

vehicle, not all of them are compatible in retarding processes or mechanisms

because the operating principle of one retarder may conict with another.

For example, an exhaust brake and a compression brake should not be used

simultaneously at high engine speeds in a turbocharged engine because the

exhaust brake operation tends to reduce the turbine pressure ratio and the

turbine speed, while a compression brake relies on high turbocharger speed

in order to deliver high boost pressure to enhance the compression-release

effect. However, for naturally aspirated engines a combination of an exhaust

brake and a compression brake has proven effective (Schmitz et al., 1992,

1994). Another example is that if the cylinder charge is basically completely

blown off during the expansion stroke in a compression brake, there will be

little air to be compressed to raise the cylinder pressure during the exhaust

stroke. Therefore, the use of an exhaust brake may not be effective in

conjunction with such a compression brake. SAE J1621 (2005) and J2458

(1998) provide information on dynamometer testing and capability rating

of engine brakes and exhaust brakes.

6.1.6 Comparison between engine brakes and drivetrain

retarders

Retarder performance is measured by the following characteristics: retarder

energy conversion ratios (related to thermodynamics), heat rejection (i.e.,

heat absorbed by the retarder and transferred to the vehicle cooling system),

maximum power, retarding power ratio (the ratio of the maximum retarding

power to the maximum ring power), low-speed torque (essentially the shape

of ‘torque vs. speed’ curve), weight, power-to-weight ratio (indicating packing

space and the stress level), power density (i.e., specic power or the ratio

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411Engine brake performance in diesel engine system design

of power to engine displacement), transient response, and noise (related to

operating mechanism). Other design attributes include durability, packaging

(installability), and cost.

In general, a hydrodynamic retarder may be able to develop a retarding

power as high as double that of the engine ring rated power. However, its

heat rejection is also very high. The cooling system used to dissipate the heat

is a great design challenge. A compression brake may be able to develop

a similar amount of retarding power as the engine ring rated power. An

exhaust brake usually produces signicantly lower retarding power. With

improved braking mechanisms and stronger engine structures, more powerful

compression brakes are being developed.

High retarding torque in the mid-speed range is very important. The torque

capability of the hydraulic and electric retarders at medium to high speeds

is usually limited by the temperature of the retarder’s cooling medium. The

drivetrain retarders often deliver poor performance (low power) at low vehicle

speeds but good performance at higher vehicle speeds, while the engine

brakes may deliver high power at low vehicle speeds. Engine brakes can

achieve this by downshifting the transmission gear to run the engine at high

speeds so that they can deliver high power on a much steeper downgrade

compared to a drivetrain retarder. The low-speed torque of an engine brake

can be increased by using an exhaust brake that throttles the turbine outlet

or inlet.

The retarder power capability or retarder design limit is often translated into

an allowable road grade for downgrade performance, or a vehicle deceleration

in a fraction of the acceleration due to gravity for level-ground performance.

For instance, a typical European legal requirement for downhill braking is

that the retarder needs to maintain 30 km/hour (18.6 mph) vehicle speed on

a 7% downgrade for 6 km (3.7 miles). Customers can be more demanding

on these requirements, for example, requiring 80 km/hr speed on a 5%

downgrade. Limpert (1975) reported that the mean vehicle deceleration when

using the engine alone was approximately 0.015 g; and using the exhaust

brake increased the mean vehicle deceleration to nearly 0.03 g. A 0.05 g

(0.5 m/s

) deceleration was reported in the literature for braking from 80

to 60 km/hr for a 40 ton truck (Schreck et al., 1992). As a comparison, a

0.15 g deceleration (1.47 m/s

) is typical in many normal vehicle stops using

the service brakes (Spurlin and Trotter, 1982), as seen in typical transit buses

from normal operating speeds to low speeds (Klemen et al., 1989).

Hydrodynamic and electromagnetic retarders are usually heavy, large and

expensive. A conventional compression brake must be structurally strong/

heavy enough to support the high loading due to the in-cylinder pressure

near the TDC (top dead center) transmitted via the braking valve, while a

compression brake actuated by variable valve timing devices can be much

lighter (e.g., reduced weight by half; Hu et al., 1997b). The exhaust-pulse

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412 Diesel engine system design

compression brake can be even lighter. The exhaust brake has very low

weight (only several kilograms) and is usually less expensive than other

retarders.

Typical engines have 5–9 hp/liter power density in motoring at 2000

rpm (SAE J1489, 2000). Exhaust brakes usually have around 10–15 hp/liter

retarding power density at 2000 rpm (Fig. 6.1). Note that the retarding power

data in this chapter are all presented as positive values for convenience.

Compression brakes may have at least 30–50 hp/liter power density (Fig.

6.2). Compression brakes usually offer higher retarding power than exhaust

brakes due to more effective retarding mechanisms of the compression-release

process during the compression and expansion strokes. Figure 6.2 shows

the retarding power of 9–16 L diesel engines equipped with conventional

compression brakes. In 1994, the highest available retarding power was 343

kW (460 hp) at 2100 rpm for the 12–15 L engine market (Freiburg, 1994).

There is a growing trend in retarding power density over the years through

design improvement. Note that the ring power density is typically about

40 hp/liter at the rated speed (e.g., 1600–2300 rpm) for big bore engines.

The retarding power density is denitely short around 1600 rpm but may

become capable at 2100 rpm. As the engine rated speed is reduced, there is an

increasingly large mismatch between the required ring speed and retarding

speed. Engine system design needs to address such a mismatch. Finally, Fig.

6.3 shows there is a large variation in the engine retarding power at a xed

engine displacement due to the variation in brake design details.

The power-to-weight ratio indicates the compactness of the retarder. The

power density of the hydrodynamic retarder is in the order of 17 hp/kg

6.1 Conventional exhaust brake performance.

0 500 1000 1500 2000 2500 3000 3500

Engine speed (rpm)

Speciﬁc retarding power (hp/liter)

The range of retarding power

of various exhaust brakes

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413Engine brake performance in diesel engine system design

(e.g., Voith Turbo’s Aquatarder). The power density of the exhaust brake is

around 30 hp/kg at 2000 rpm engine speed. The power-to-weight ratio of

the conventional compression brake is around 10–11 hp/kg.

The friction brake usually has very fast transient response and high torque

at low speed. The hydrodynamic retarder has a lling time and a delay in

response, especially for a large-volume retarder. A small torus volume can

improve the response time. The electromagnetic retarder usually has very

0 500 1000 1500 2000 2500

Engine speed (rpm)

Speciﬁc retarding power (hp/liter)

6.3 Competitive benchmarking analysis on exhaust brake retarding

power.

5.0 5.5 6.0 6.5 7.0 7.5 8.0 8.5 9.0

Engine displacement (liter)

1 hp = 0.7457 kW

Note: The retarding

power is plotted as

positive values here.

Competitive engine exhaust brake

performance comparison at 2750 rpm

Engine braking retarding power (hp)

240

230

220

210

200

190

180

170

160

150

140

130

120

6.2 Conventional compression-release brake performance.

The range of retarding power

of various compression brakes

2000s

1990s

1980s

1970s

1960s

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414 Diesel engine system design

fast response time. The transient response sensitivity of the hydraulic system

in the conventional compression brake can be fairly rapid (Greathouse

et al., 1971). The compression brake can achieve full operating status in about

0.2 s and deactivate in less than 0.1 s, as indicated by Morse and Rife (1979).

However, turbocharger lag may affect the transient response in terms of the

intake manifold boost pressure for the compression brake in turbocharged

engines.

Drivetrain retarders and the exhaust brake are very quiet (usually no

louder than a plain engine). The conventional compression brake is noisy

due to the high-frequency exhaust pressure pulse waves caused by the sudden

compression-release process. The bleeder brake is less noisy. Engine brake

noise is discussed in more details in Section 6.4.8.

6.1.7 Retarder energy conversion ratios

In drivetrain retarders, all the mechanical power is basically converted to

frictional heat and there is a challenge to manage this heat transfer rate. There

are two methods to manage the heat: using a waste heat recovery device, or

using a cooling system to dissipate the heat to the ambient.

For engine brakes, their energy conversion needs to be analyzed based

on the rst law of thermodynamics for the engine system. It is well known

that the fuel energy basically goes into three parts in engine ring operation

based on energy balance: the brake (shaft) work, the exhaust enthalpy, and

the coolant heat. Similarly, in engine brake operation, when the fuel energy

is zero, according to the engine energy balance the mechanical brake power

(i.e., retarding power) needs to be balanced by the exhaust enthalpy rate

and the heat rejection to the coolant. In other words, because the engine is

essentially an energy conversion device (e.g., as an air compressor in either

the compression brake or the exhaust brake), the higher the retarding power

absorbed by the engine, the higher the exhaust enthalpy rate and the heat

rejection.

The energy conversion of the retarders needs to be evaluated and compared

for their effects on engine performance. The two subjects of concern are: (1)

the heat rejection and the associated difculty for the cooling system design;

(2) the exhaust gas enthalpy and its impact on aftertreatment operation (e.g.,

in regeneration) and turbocharger transient operation. It should be noted that

the exergy of the coolant heat is so low that it is usually not worth the effort

of waste heat recovery.

In order to compare different retarders, two retarder energy conversion

ratios are dened as follows. The retarder heat dissipation ratio is dened

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415Engine brake performance in diesel engine system design

hrjrhr

etarding

hrjhr

firin

rh,rh

jr,jr

firin,,firin

gm,,gm





6.8

where



hrjrhr

etarding

jr,jr

is the unrecoverable heat dissipation to the cooling system

during the retarder operation, and



hrjhr

firin

j,,j

firin,,firin

gm,,gm

is the maximum engine coolant

heat rejection during the  ring operation.

The retarder available exhaust energy ratio is de ned as

exh

reta

exh

firin

firin,,firin

gm,,gm





6.9

where



exh

reta

is the exhaust enthalpy  ow rate during the retarder

operation, and



exh

firin

firin,,firin

gm,,gm

is the maximum exhaust enthalpy  ow rate

during the  ring operation.

Obviously, a lower retarder heat dissipation ratio less than 1 and a high

retarder available exhaust energy ratio are desirable for any retarders. The

r,heat

of drivetrain retarders is usually close to or much greater than 1, which

imposes a signi cant design challenge for the vehicle cooling system. The

r,heat

of engine brakes is usually much less than 1. Only when the engine

brake’s retarding power is several times higher than the engine rated power

is its heat dissipation ratio close to 1.

The h

r,exh

of drivetrain retarders is very low because only the exhaust heat

from engine motoring contributes to the h

r,exh

. This means during a long

downgrade the exhaust temperature available for engine aftertreatment and

turbocharger operations is very low. In contrast, the h

r,exh

of engine brakes

is usually greater than or close to 1 at the full retarding power. The exhaust

 ow in engine braking contains a signi cant amount of energy to keep the

aftertreatment and the turbocharger at good operating conditions.

Therefore, it is clear that engine brakes are far superior to drivetrain

retarders from a thermodynamic perspective. This is especially true if the

comparison between the retarders is examined from the second law of

thermodynamics. Drivetrain retarders convert all the mechanical energy

of the vehicle motion to frictional heat, which has a very low availability

or exergy. Engine brakes convert the mechanical energy of the vehicle to

useable exhaust energy, which has a high availability.

6.1.8 Retarding process efﬁ ciency of compression

brakes

The research on the retarding process ef ciency of compression brakes

was started by Israel and Hu (1993) when they analyzed the performance

sensitivity of the Jacobs compression brake to the braking system compliance

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416 Diesel engine system design

at different ambient conditions. Later, Hu et al. (1997a) applied retarding

process efciency to the study of a VVA compression brake to try to quantify

its performance benet and compare it with the conventional compression

brake. They pointed out that the maximum retarding power available comes

from an isentropic compression followed by an instantaneous sudden release

of the compressed cylinder charge at the TDC. In a real system, inevitable

losses are caused by all the factors contributing to the insufcient compression

and the non-instantaneous release of the cylinder charge (Fig. 6.4). Some

examples are as follows:

∑ There are heat losses during the compression stroke.

∑ There is a leakage or insufcient compression near the end of the

compression process caused by the opening of the braking valve. The

braking valve needs to be opened before the TDC because it cannot

open instantaneously with a sufciently large ow area right at the TDC.

Note that there is a restriction of valve-to-piston clearance at the TDC.

Moreover, the braking valve opening timing is selected to try to limit

the maximum cylinder pressure and the associated gas loading acting

on the braking assembly from a structural durability point of view.

∑ The residue cylinder pressure during the expansion stroke caused by

incomplete blow-down result in some positive power. This is caused

mainly by the restriction of the supersonic ow across the braking

valve.

∑ The hydraulic compliance of the engine brake system may lower the

braking valve lift and increase the residue pressure in the expansion

stroke.

Loss due to early blow-down

Loss due to incomplete blow-down and

residue pressure producing positive power

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1

Instantaneous in-cylinder volume

Cylinder pressure (bar, absolute)

6.4 Illustration of retarding process efﬁciency of compression brake.

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417Engine brake performance in diesel engine system design

Israel and Hu (1993) de ned the retarding process ef ciency of compression

brakes as the ratio of the actual retarding power to the isentropic retarding

power during the compression and expansion strokes, given by

rCRrC

ctual

ntr

opi

rC,rC





ra,ra

ri,ri

6.10

The pumping loss and mechanical friction are excluded in the ef ciency

de nition of equation 6.10. In order to calculate the retarding process

ef ciency, the actual retarding power can be computed based on the cylinder

p–V diagram or can be measured in the compression and expansion strokes.

Israel and Hu (1993) proposed that the isentropic retarding power can be

computed by the following equation during the compression stroke, assuming

the power is zero during the expansion stroke:



ntr

opi

ri,ri

(

)

(

pC,pC

(

)

)1 )

11 +11

6.11

where p

is the compressor inlet pressure, p

is the compressor outlet pressure,

is the compressor inlet temperature, h

is the compressor ef ciency,

and W is the engine compression ratio. It should be noted that equation

6.11 is only valid for a simpli ed closed thermodynamic system under the

assumption that all the engine valves are closed from the BDC to TDC for

an isentropic compression. In reality, the intake valve actually closes about

40∞ crank angle after the BDC. Moreover, equation 6.11 is not applicable

for the case of having braking gas recirculation (BGR, to be detailed later)

where the exhaust valve is open during the compression stroke and the initial

state of the gas at the beginning of the compression is complicated by the

BGR gas mixing.

The above retarding process ef ciency for the compression brakes is a

measure of the perfection of the thermodynamic process. The de nition

provides a fair basis to compare the performance of a given compression

brake at different design or operating conditions (e.g., engine speeds, ambient

conditions), or to compare the effectiveness of different types of compression

brakes. In order to increase the actual retarding power, either the retarding

process ef ciency or the ideal isentropic maximum available power needs

to be increased. In order to increase the isentropic power, intake manifold

pressure or engine compression ratio needs to be increased. This will result

in an increase in peak cylinder pressure, which in fact is limited by the

maximum gas loading that can be sustained by the braking assembly.

It is dif cult to de ne a similar ef ciency for the retarding process in

the pumping loss strokes. Recall from Chapter 4 that the pumping loss is

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