Xin Q. Diesel Engine System Design

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820 Diesel engine system design

Siano D and Bozza F (2009), ‘Combustion noise prediction in a small diesel engine nalized

to the optimization of the fuel injection strategy’, SAE paper 2009-01-2077.

Tabuchi T, Fujitani N and Makino N (1995), ‘Recent technology to improve engine

combustion noise and exhaust emission by optimizing fuel injection system’, SAE

paper 951791.

Torregrosa A J, Broatch A, Martin J and Monelletta L (2007), ‘Combustion noise

level assessment in direct injection diesel engines by means of in-cylinder pressure

components’, Measurement Science and Technology, 18, 2131–2142.

Tung V T C and Crocker M J (1982), ‘Diesel engine noise and relationship to cylinder

pressure’, SAE paper 820237.

Wang S, Chalu C and Gautier F (2007), ‘Optimization of combustion noise of modern

diesel engines for passenger cars’, SAE paper 2007-01-2379.

Wolschendorf J, Dürnholz M and Schmillen K (1991), ‘The IDI diesel engine and its

combustion noise variations’, SAE paper 910228.

11.13.3 References and bibliography on piston slap noise

Chien M H (1995), ‘Engine impact measurement and quantication’, SAE paper

951236.

DeJong R G and Parsons N E (1982), ‘Piston slap noise reduction in a Vee-block diesel

engine’, SAE paper 820240.

De Luca J C and Gerges S N Y (1996), ‘Piston slap excitation: literature review’, SAE

paper 962395.

Fabi P, Flekiewicz M, Madej H, Wojnar G and Flekiewicz B (2007), ‘Inuence of piston

slap on engine block vibration’, SAE paper 2007-01-2163.

Furuhama S and Hirukawa K (1983), ‘Piston slap motion and engine noise under low

temperature idling operation of diesel engines’, SAE paper 830066.

Gerges S N Y, De Luca J C and Lalor N (2005), ‘Effect of cylinder lubrication on piston

slap’, SAE paper 2005-01-2165.

Haddad S D (1995), ‘Theoretical treatment of piston motion in IC piston engine for

the prediction of piston slap excitation’, Mechanism and Machine Theory, 30(2),

253–269.

Haddad S D and Howard D A (1980), ‘Analysis of piston slap-induced noise and assessment

of some methods of control in diesel engines’, SAE paper 800517.

Haddad S D and Tjan K-T (1995), ‘An analytical study of offset piston and crankshaft

designs and the effect of oil lm on piston slap excitation in a diesel engine’, Mechanism

and Machine Theory, 30(2), 271–284.

Kageyama H, Hara S and Kawabata Y (1994), ‘Study of the simulation of piston skirt

contact’, JSAE paper 9430095.

Kaiser H J, Schmillen K and Spessert B (1988), ‘Acoustical optimization of the piston

slap by combination of computing and experiments’, SAE paper 880100.

Kamiya M, Kobayashi T, Mihara Y and Someya T (2007), ‘Measurement of piston skirt

oil-lm pressure under piston slap’, SAE paper 2007-01-2215.

Kamp H and Spermann J (1995), ‘New methods of evaluating and improving piston

related noise in internal combustion engines’, SAE paper 951238.

Kobayashi T, Takahashi Y and Bell D J (2007), ‘How to predict the piston slap-noise

using 3D piston motion simulation’, SAE paper 2007-01-1245.

Künzel R, Tunsch M and Werkmann M (2000), ‘Piston related noise with spark ignition

Diesel-Xin-11.indd 820 5/5/11 12:01:49 PM

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821Noise, vibration, and harshness (NVH) in engine system design

engines – characterization, quantication and mechanisms of excitation’, SAE paper

2000-01-3311.

Künzel R, Werkmann M and Tunsch M (2001), ‘Piston related noise with diesel engines

parameters of inuence and optimization’, SAE paper 2001-01-3335.

Lalor N, Grover E C and Priede T (1980), ‘Engine noise due to mechanical impacts at

pistons and bearings’, SAE paper 800402.

Munro R and Parker A (1975), ‘Transverse movement analysis and its inuence on diesel

piston design’, SAE paper 750800.

Nakada T, Yamomoto A and Abe T (1997), ‘A numerical approach for piston secondary

motion analysis and its application to the piston related noise’, SAE paper 972043.

Oetting H, Pundt D and Ebbinghaus W (1984), ‘Friction in the piston group and new

ideas for piston design’, SAE paper 841299.

Ohta K, Irie Y, Yamamoto K and Ishikawa H (1987), ‘Piston slap induced noise and

vibration of internal combustion engines (1st report, theoretical analysis and simulation)’,

SAE paper 870990.

Paranjpe R S (1998), ‘Development of a math-based piston noise model’, SAE paper

980564.

Patel P, Mourelatos Z P and Shah P (2007), ‘A comprehensive method for piston secondary

dynamics and piston-bore contact’, SAE paper 2007-01-1249.

Richmond J W and Parker D A (1987), ‘The quantication and reduction of piston slap

noise’, Proc. IMechE, Part D: Transport Engineering, 201(4), 235–244.

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paper 65-OGP-10.

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(1994), ‘Engine experiments on the effects of design and operational parameters on

piston secondary motion and piston slap’, SAE paper 940695.

Sander W, Steidle W and Wacker E (1979), ‘Piston movement and its inuence on noise

of automotive engines’, SAE paper 790272.

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Engine Noise – Excitation, Vibration, and Radiation, Plenum Press, New York.

Teraguchi S, Suzuki W, Takiguchi M and Sato D (2001), ‘Effects of lubricating oil

supply on reductions of piston slap vibration and piston friction’, SAE paper 2001-

01-0566.

Usami T, Wada S and Sonoda S (1975), ‘Piston slap noise of indirect combustion diesel

engine’, SAE paper 750801.

Vora K C and Ghosh B (1991), ‘Vibration due to piston slap and combustion in gasoline

and diesel engines’, SAE paper 911060.

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SAE paper 901491.

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in the sliding bearing’, Mechanism and Machine Theory, 9, 61–80.

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‘A numerical model of piston secondary motion and piston slap in partially ooded

elastohydrodynamic skirt lubrication’, SAE paper 940696.

Xin Q (1999), Analytical and Numerical Study of Lubrication Dynamics for Internal

Combustion Engine Piston Assemblies and Journal Bearings, D.Sc. (Doctor of Science)

degree dissertation, Washington University, St Louis, Missouri, USA.

Yonezawa T and Kanda H (1985), ‘Analysis of cavitation erosion on cylinder liner and

cylinder block’, SAE paper 850401.

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825

Diesel engine heat rejection and cooling

Abstract: This chapter develops the analysis methods of heat rejection

based on the engine energy balance of the rst law of thermodynamics for

diesel engine system design. A theoretical analysis of engine miscellaneous

heat losses forms the foundation of the analysis approach. The concept of

base engine coolant heat rejection is proposed and its characteristics are

illustrated. Cooling system design calculations are then discussed to address

cooler performance, cooling capability, and coolant temperatures.

Key words: heat rejection, cooling, rst law of thermodynamics, energy

balance, miscellaneous heat losses, base engine coolant heat rejection, low-

heat-rejection engine, cooler, sink temperature, waste heat recovery.

12.1 Engine energy balance analysis

12.1.1 Challenges from thermal load to heat rejection

control

Turbocharged diesel engines have a higher thermal load than non-turbocharged

diesel engines. Engine heat transfer and thermal loading have been extensively

researched in the past several decades. Engine heat transfer fundamentals

are introduced by Heywood (1988). Thermal loading and engine cooling are

discussed in detail by Sitkei (1974) and Challen and Baranescu (1999). The

thermal load refers to the gas temperature and the heat ux a component is

exposed to. A higher thermal load in an engine is typically characterized by

higher gas temperatures in the intake manifold (e.g., caused by the hot air from

the compressor outlet), the cylinder, and the exhaust manifold. The higher

thermal load causes many durability problems, such as cracks in the cylinder

head and exhaust manifold, and damage in the turbine rotor. Managing the

thermal load and reducing the exhaust manifold gas temperature have been

important design challenges for turbocharged engines.

The successful measures in thermal load control for non-EGR engines used

to include the following: (1) reducing the intake manifold temperature by

using air charge cooling; (2) reducing the compressor outlet air temperature

by increasing the compressor efciency; (3) increasing fresh air scavenging

from the intake port to the exhaust port by increasing the pressure differential

between the intake manifold and the exhaust manifold through turbocharger

matching or manifold design; (4) increasing fresh air scavenging by increasing

the valve overlap through cam design; (5) increasing the air–fuel ratio to

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826 Diesel engine system design

reduce the exhaust temperature; (6) enhancing the cooling capacity of the

cylinder head; and (7) adjusting the fuel injection timing and the shape of

heat release rate to reduce the exhaust temperature.

However, many of the above measures cannot be applied to modern

EGR engines. For example, enhancing fresh air scavenging with a negative

engine delta P (i.e., intake manifold pressure higher than exhaust manifold

pressure) is impossible due to the need to drive EGR ow. On the other

hand, for EGR engines adding cooled EGR can reduce the in-cylinder gas

temperature and the NO

concentration due to the higher specic heat of the

EGR gases. Meanwhile, the exhaust gas temperature and the thermal load

can be reduced accordingly. However, cooling the EGR increases the total

engine heat rejection (heat transfer rate) regardless of whether it is cooled

by engine coolant, cooling air or low-temperature radiator coolant. The

heat rejection increases as more EGR is used. Essentially, in modern diesel

engines, a large portion of the thermal load is moved by EGR from the hot

gas to the coolant because the engine needs to run cooler in order to meet

the NO

emissions requirement. The design challenge is then shifted to how

to manage the heat rejection due to the EGR gas. In a modern powertrain

development program, one of the top priorities is to control the engine heat

rejection and design a packageable cooling system to control the radiator

inlet coolant temperature.

The coolant heat rejection in modern diesel engines increases mainly for

three reasons: high EGR rate, low intake manifold gas temperature, and high

power rating. In order to meet stringent emissions regulations, high EGR

rate and very cold EGR are required to achieve low oxygen concentration

and low mixture gas temperature in the intake manifold. The increasing

customer demand on engine power rating results in a proportional increase

in the fueling rate and the coolant heat rejection. Owing to the difculties in

heat rejection control and cooling system design (Bowman et al., 2005), all

the heat rejection data need to be scrutinized for accuracy. Any inaccurate

estimation method becomes obsolete. A critical task is to predict heat

rejections (including the important cylinder heat rejection) accurately at

various extreme ambient conditions.

12.1.2 System methodology of heat rejection and cooling

analysis

Experimental errors in coolant ow rate and temperature cause variations

in the measurement of engine coolant heat rejection. The heat rejection

errors can range from 3 to 20 kW and become unacceptably large for precise

design. In order to measure the engine cylinder coolant heat rejection directly,

experimental errors need to be minimized. Alternatively, an indirect approach

with sufcient accuracy can be used to calculate the coolant heat rejection.

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827Diesel engine heat rejection and cooling

A numerical simulation of the cylinder heat transfer by using in-cylinder

CFD is usually inaccurate or impractical due to the extremely complicated

conditions of the in-cylinder turbulent  ows. A practical approach needs to be

developed for engine system design to handle the prediction of coolant heat

rejection and cooler design parameters. This chapter presents a methodology

for the calculations. The methodology includes the following features:

1. The foundation is built upon estimating engine miscellaneous heat

losses.

2. Two energy balance methods of the  rst law of thermodynamics with

different boundaries are used to calculate and compare the cylinder heat

rejection.

3. The percentage of base engine coolant heat rejection lost to fuel energy

is selected as a characteristic parameter for the basis of heat rejection

analysis.

4. The Woschni correlation of in-cylinder heat transfer is selected in an

engine cycle simulation model and is calibrated on the basis of item (3)

above.

5. The cooling medium temperatures of the coolers and the radiator are

estimated.

6. The cooling capability of the coolers and the radiator is de ned.

12.1.3 Energy balance analysis methods of the ﬁ rst law

of thermodynamics

As a part of the methodology of heat rejection analysis in engine system

design, engine energy balance equations are developed to use the gas  ow

rates and gas temperatures to calculate heat rejections. Two methods of

calculating the energy balance are presented below, depending on where

the control volume boundary is drawn for the system.

Method 1 sets the control volume boundary for gas  ows at the compressor

inlet and the turbine outlet, and the boundary for the fuel  ow is set at the

fuel tank. If low-pressure-loop EGR is used, the boundary is set after the

EGR pickup. For a coolant-cooled turbocharged engine, as an example, the

energy balance equation is given as follows:







comp

fuel

Fuel

FuelInFuel

fuel

LHV





+WW +



Eacc

exh

bas

e–

coolant

++ +



EGRcooler

12.1



++QQ++

+QQ +

++QQ++

CISCCI

FuelCooler

miscell

aneous

miscellaneousmiscell

111

In equation 12.1,



comp

is the compressor air or gas  ow rate, h

CompIn

the speci c enthalpy of the compressor inlet air or gas  ow,



fuel

is the

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828 Diesel engine system design

fuel  ow rate, and h

FuelIn

is the speci c enthalpy of the fuel  ow at the fuel

tank temperature. Note that



fuel

Fuel

FuelInFuel

is usually very small compared to

the enthalpy of the air  ows. The q

LHV

is the lower heating value of the fuel

which is related to fuel formulation (see ASTM American National Standard

D4868-00 and SAE J1498. Note that the enthalpy of vaporization of diesel

fuels is usually small compared to their lower heating value.



is the engine

 ring brake power.



Eacc

is a certain engine accessory power including the

alternator, the air compressor, and the cooling fan. The



exh

is the exhaust

gas  ow rate, and h

TurbOut

is the speci c enthalpy of the turbine outlet exhaust

gas  ow.



bas

e–

coolant

is de ned as the base engine coolant heat rejection (to

be detailed later).



EGRcooler

is the EGR cooler heat rejection.



is the

charge air cooler heat rejection.



is the compressor inter-stage cooler

heat rejection (if any).



FuelCooler

is the fuel cooler heat rejection (if any).



miscell

aneous

miscellaneousmiscell

is the miscellaneous heat losses such as the convection and

radiation heat transfer from the exhaust manifold, the EGR tubing (excluding

the EGR cooler), the engine block and the turbocharger surface, and the

thermal energy of the unburned or incompletely burned fuel (up to 1–2%

of the fuel energy).

Method 2 sets the control volume boundary for gas  ows at the intake

manifold and the turbine inlet, and the boundary for the fuel  ow is set at

the fuel tank. If high-pressure-loop EGR is used, the boundary is set before

the EGR pickup. As an example, for a coolant-cooled turbocharged engine

the energy balance is given as follows:



rIMTrI

MT,aMT

EGRIEG

RIMTRI

MT,EMT

fuel

Fuel

FuelInFuel

fuel

ffuelf

LHV





+WW +

Eacc

EGRpum

+WW +

EGRpumpsEGRpum

EGRpum

upe

har

hargehar

har

eexe

bInbI

urTur

12.2





+QQ +

bas

e–

coolant

FuelCooler

miscell

aneou

miscellaneoumiscell

aneouusaneou

In equation 12.2, h

IMT,air

is the speci c enthalpy of the air  ow entering the

intake manifold, h

IMT,EGR

is the speci c enthalpy of the EGR  ow entering

the intake manifold, h

TurbIn

is the speci c enthalpy of the turbine inlet exhaust

gas  ow,



EGRpum

EGRpumpEGRpum

is the EGR pump power (if any), and



upe

har

hargehar

is the

mechanical supercharger power (if any). It should be noted that



miscell

aneous

miscellaneousmiscell

is lower than



miscell

aneous

miscellaneousmiscell

. Moreover, the heat rejections of the transmission

cooler and other small auxiliary coolers are not explicitly included in the

above equations. For example, the heat generated inside the transmission

comes from the gear or torque converter frictional power losses, which are

already a part of the engine crankshaft output power. Similarly, vehicle

accessory resistance power, such as the power from the power steering pump

and the air conditioning compressor, needs to be overcome by a part of the

engine crankshaft power



Given the turbocharger power balance between the compressor and the

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829Diesel engine heat rejection and cooling

turbine, equations 12.1 and 12.2 can be derived as equivalent (i.e., if equation

12.1 is subtracted from 12.2, all the terms will cancel out to zero on both

sides of the equality sign). The gas  ow enthalpy is calculated by using

hc = hc

· d

where the speci c heat c

must be used as a function of temperature and

gas species composition for acceptable accuracy. The cooler heat rejection

can be calculated by the following:



cooler

gas

ooler

oolerInooler

letGas

Cooler

Qm =Qm

(

Qm (Qm



(



cT (cT

gas

(

gas

(

cT (cT

–

letGas

lletGasl

)

12.3

In order to characterize a base engine heat rejection, it is important to

distinguish two types of working  uids: (1) the in-cylinder  ows (including

air, EGR, and fuel); and (2) the outside-cylinder cooling media (including

coolant and oil  ows). The pressures and temperatures of the gas and fuel

 ows at the entry of the cylinder are affected by their respective pumping

power (e.g., a compressor or supercharger for the fresh air  ow, an EGR

pump for the EGR gas  ow, a fuel pump for the fuel  ow) and cooler heat

rejection (e.g., charge air cooler, EGR cooler, fuel cooler). The in-cylinder

gas  ows participate in the engine cycle processes and their energy leaves

the cylinder as shaft work, exhaust enthalpy, and heat rejection. On the

other hand, the water pump power and the oil pump power for pumping the

outside-cylinder cooling media eventually only dissipate as heat rejection.

Therefore, it is usually convenient to de ne the ‘base engine coolant heat

rejection’ to include the heat rejections from the following:

∑ the cylinders (the piston, the valves, the liner, and the cylinder head,

including the exhaust ports)

∑ the oil cooler (including piston cooling and mechanical rubbing

friction)

∑ the water pump power and the oil pump power.

The base engine heat rejection is a critical parameter which characterizes the

fundamental design and operating features of the base engine (i.e., excluding

the effect of the coolers, except for the oil cooler). Different engines can

be compared on the common basis of the percentage of fuel energy lost to

the base engine coolant heat rejection in order to evaluate their performance

competitiveness. It should be noted that if the cylinder heat transfer is the

focus of the study, engine friction and water/oil pump power should be

excluded from the de nition of base engine coolant heat rejection.

In equations 12.1 and 12.2, the only unknown is



bas

e–

coolant

If there were

no experimental errors in gas temperature and  ow rate measurements, the

same result of base engine coolant heat rejection would have been obtained

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830 Diesel engine system design

by using either method 1 or 2. Therefore, the above energy balance methods

can serve the following ve purposes:

1. Calculate coolant heat rejection accurately for cooling system design,

including the cylinder heat rejection. Note that the vehicle radiator coolant

heat rejection is equal to the sum of the heat rejections from the base

engine, EGR cooler, and other coolant-cooled coolers.

2. Compare the calculated coolant heat rejections from methods 1 and 2

with available heat rejection data coming from the direct coolant ow

measurements in order to verify the data accuracy for each method.

3. Check the accuracy of engine performance test data of gas temperatures

and ow rates.

4. Reveal the engine energy balance distribution based on the rst law

of thermodynamics, and reveal the competitiveness of the base engine

design in term of heat losses.

5. Tune the cylinder heat transfer model (e.g., a multiplier for the Woschni

in-cylinder gas-side heat transfer coefcient) in the engine cycle simulation

to match with the engine test data on the percentage of fuel energy lost

to the base engine coolant heat rejection.

Figure 12.1 shows a tuned in-cylinder heat transfer coefcient. Note that

the heat transfer coefcient is at the gas side from the in-cylinder gas to the

cylinder inner wall. It is much more dominant in terms of thermal resistance

than the coolant-side heat transfer coefcient if compared in the overall

heat transfer coefcient from the gas to the coolant. Also note that when

Cylinder pressure (psi)

Cylinder gas temperature (K)

In-cylinder heat transfer coefﬁcient

(W/m

.K)

1 psi = 0.06895 bar

5000

4500

4000

3500

3000

2500

2000

1500

1000

500

–90 0 90 180 270 360 450 540 630

TDC Crank angle (degree)

12.1 Engine in-cylinder cycle simulation of gas pressure,

temperature, and heat transfer.

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831Diesel engine heat rejection and cooling

the fuel injection timing is advanced, the shape of the in-cylinder pressure,

temperature, and heat transfer coefcient traces change or shift accordingly,

and the percentage of base engine coolant heat rejection in fuel energy

increases. Figure 12.2 shows an example of the energy balance calculations

using engine test data at full load. It is observed that the results given by

the two methods match well. Figure 12.3 shows typical energy balance

distributions of diesel engines. Another example of diesel engine energy

balance was provided by Wallace and Kremer (2009).

12.2 Engine miscellaneous energy losses

12.2.1 Overview of miscellaneous energy losses

In the energy balance equations 12.1 and 12.2, the miscellaneous loss is

treated as a known quantity in order to calculate the base engine coolant heat

rejection. The miscellaneous loss is not negligible for the precise control of

heat rejections in modern high-EGR engines (Fig. 12.3). In fact, the analysis

of the miscellaneous loss is very complex. Its heat transfer part is related to

engine speed (i.e., the time scale for heat transfer) and the exhaust manifold

gas temperature that depends on engine load. Only after the miscellaneous

heat losses are calculated, can the coolant heat rejection be estimated in

engine system design by using method 1 or 2. Therefore, the analysis of the

miscellaneous heat losses is of paramount importance in the heat transfer

theory in diesel engine system design. The complete theory on miscellaneous

heat losses was developed by Xin and Zheng (2009).

Xin and Zheng (2009) found that the percentage of fuel energy lost to the

Method 1 (control

volume boundary set

at compressor inlet

and turbine outlet)

Method 2 (control

volume boundary set

at intake manifold

and turbine inlet

before EGR tapping)

Peak torque

Rated power

Engine speed

Percentage of fuel energy lost to base

engine coolant heat rejection

(excluding EGR cooler heat rejection)

35%

30%

25%

20%

15%

12.2 Energy balance calculation on full-load lug curve.

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