Xin Q. Diesel Engine System Design
Подождите немного. Документ загружается.
© Woodhead Publishing Limited, 2011
F: percentage of
fuel energy lost to
miscellaneous losses
E: percentage of fuel
energy converted to
brake power
D: percentage of
fuel energy lost to
sensible exhaust
gas energy at the
turbine outlet
C: percentage of
fuel energy lost to
charge air cooler
heat rejection
B: percentage of fuel
energy lost to EGR
coolant heat rejection
A: percentage of fuel
energy lost to base
engine coolant heat
rejection (excluding EGR
cooler heat rejection)
US EPA 2004 emissions HD diesel engine energy balance at full load
A
A + B
A + B + C
A + B + C
+ D
A + B + C +
D + E
A + B + C +
D + E + F
500 1000 1500 2000 2500 3000 3500 4000
Engine speed (rpm)
100%
90%
80%
70%
60%
50%
40%
30%
20%
10%
0%
Percentage of fuel energy
12.3 Engine energy balance comparison between non-EGR and EGR engines.
Diesel-Xin-12.indd 832 5/5/11 12:02:50 PM
© Woodhead Publishing Limited, 2011
Energy balance of non-EGR high NO
x
engine Energy balance of EGR low NO
x
engine
Miscellaneous
heat losses,
3.00%
Miscellaneous
heat losses,
2.83%
Turbine outlet
exhaust
enthalpy, 32%
Turbine outlet
exhaust
enthalpy, 20%
EGR cooler
heat rejection,
11.82%
Charge air
cooler heat
rejection,
9.00%
Charge air
cooler heat
rejection,
10.35%
Base engine
coolant heat
rejection, 17%
Base engine
coolant heat
rejection, 16%
Engine brake
power, 39%
Engine brake
power, 39%
(EPA 2007-2010 emissions)
12.3 Continued
Diesel-Xin-12.indd 833 5/5/11 12:02:50 PM
834 Diesel engine system design
© Woodhead Publishing Limited, 2011
miscellaneous heat losses is a function of certain characteristic gas temperature
(e.g., exhaust manifold gas temperature), cooling medium temperature, engine
speed and load. Figure 12.4 shows that the percentage of the miscellaneous
losses increases when engine speed or load decreases. They also evaluated
the miscellaneous heat losses for a water-cooled diesel engine using energy
balance method 1. If the losses due to incomplete combustion are negligible
and natural convection heat transfer occurs around the engine in a test cell
or an engine compartment, 3% of fuel energy lost to the miscellaneous heat
is a reasonable estimate with method 1 at full load from peak torque to rated
power. The calculated percentage difference between methods 1 and 2 is
about 1.8% (Fig. 12.5).
12.2.2 Parametric dependency of miscellaneous heat
losses
Denote the percentage of miscellaneous heat losses in fuel energy as
G
Q
Q
miscell
aneous
miscellaneousmiscell
fuel
=
12.4
where
Q
fuel
is the fuel energy rate. The miscellaneous heat losses of energy
balance method 1 can be expressed in the following form (Xin and Zheng,
2009):
QC
CT
CT
miscell
QC
miscell
QC
aneous
QC
aneous
QC
miscellaneousmiscell
QC
miscell
QC
aneous
QC
miscell
QC
ch
CT
ch
CT
ch
CT
ch
CT
,1
QC
,1
QC
12
CT
12
CT
3
CT
3
CT
4
QC =QC
QC QC
12
12
+
12
+
12
+
12.5
10% load
20% load
40% load
50% load
70% load
100% load
1000 1500 2000 2500 3000
Engine speed (rpm)
12
11
10
9
8
7
6
5
4
3
2
1
0
Percentage of fuel heat energy lost
to miscellaneous losses (%)
12.4 Miscellaneous heat losses as functions of engine speed and load
(calculated with method 1).
Diesel-Xin-12.indd 834 5/5/11 12:02:51 PM
835Diesel engine heat rejection and cooling
© Woodhead Publishing Limited, 2011
where C
1
, C
2
and C
3
are coef cients. T
ch
is a characteristic gas temperature (e.g.,
exhaust manifold gas temperature). The term C
2
T
ch
represents the convective
heat transfer from the engine surface to the cooling medium. The term C
3
T
ch
4
represents the radiation heat transfer. Note that the complex nonlinear effect
of the temperature in radiation heat transfer is lumped into the coef cients.
Substituting equation 12.5 into 12.4, the following is obtained:
G
CC
TC
T
CC
NC
JC
N
ch
TC
ch
TC
ch
T
ch
T
EE
NC
EE
NC
E
1
12
CC
12
CC
3
4
45
CC
45
CC
67
JC
67
JC
EE67EE
JC
EE
JC
67
JC
EE
JC
=
CC +CC
CC
12
CC +CC
12
CC
TC TC
TC
ch
TC TC
ch
TC
12
12
CC
12
CC CC
12
CC
TC+ TC
CC +CC
CC
45
CC +CC
45
CC
NC NC
NC
EE
NC NC
EE
NC
45
45
CC
45
CC CC
45
CC
NC+ NC
NC
EE
NC+ NC
EE
NC
JC +JC
JC
67
JC +JC
67
JC
JJJ
E
12.6
where the fuel energy rate is modeled as a function of engine speed N
E
and
engine brake torque J
E
according to the trend observed from engine test data,
and the subscript 1 of G represents energy balance method 1.
The characteristic temperature (exhaust manifold gas temperature here)
is affected primarily by engine speed, load, air–fuel ratio, EGR rate, and
fuel injection timing. The measured exhaust manifold gas temperature is
a function of engine speed and load. It is observed that the temperature is
predominantly linearly proportional to the engine torque within a wide range
with a weak dependency on engine speed.
When the engine torque J
E
is kept constant, the numerator of equation 12.6
can be simpli ed as a constant. Therefore, the percentage of miscellaneous
heat losses of energy balance method 1 becomes
10% load
20% load
40% load
50% load
70% load
100% load
1000 1500 2000 2500 3000
Engine speed (rpm)
4.0
3.5
3.0
2.5
2.0
1.5
1.0
The difference in percentage of miscellaneous
heat losses between methods 1 and 2 (%)
12.5 Calculated difference in the percentage of miscellaneous heat
losses between the energy balance methods 1 and 2.
Diesel-Xin-12.indd 835 5/5/11 12:02:51 PM
836 Diesel engine system design
© Woodhead Publishing Limited, 2011
G
C
CC
N
J
E
1,
8
91
CC
91
CC
0
=
CC +CC
CC
91
CC +CC
91
CC
12.7
where C
8
, C
9
and C
10
are coef cients. The parametric dependency of the
percentage of miscellaneous heat losses upon engine speed shown in Fig.
12.4 is theoretically explained by equation 12.7.
If the engine speed is constant, the percentage of miscellaneous heat losses
is
G
CC
TC
T
CC
J
CC
N
1,
12
CC
12
CC
3
4
11
CC
11
CC
12
12
CC
12
CC
=
CC +CC
CC
12
CC +CC
12
CC
TC TC
TC TC
12
12
CC
12
CC CC
12
CC
TC+ TC
CC +CC
CC +CC
CC
12
CC +CC
12
CC
TC
ch
TC
ch
TC
ch
TC
TC TC
ch
TC TC
ch
T
ch
T
E
ª
TTT
CC
J
ch
TTT
ch
TTT
E
11
CC
11
CC
12
CC +CC
12.8
where C
11
and C
12
are coef cients. Generally, the radiation term C
3
T
ch
4
in
equation 12.8 is small. For simplicity, C
3
= 0. From engine test data it was
observed that the engine brake torque increases much faster than the exhaust
temperature. For example, when the engine brake torque increases sixfold
from 100 ft.lb to 600 ft.lb, the exhaust manifold gas temperature becomes
only approximately doubled from 450°F to 950°F. Therefore, as engine brake
torque increases, the percentage of miscellaneous heat losses decreases.
12.2.3 Engine scaling for miscellaneous heat losses
The miscellaneous heat losses can be calculated by heat transfer equations once
the input data are complete. When the engine block emissivity, the block area,
the convection heat transfer coef cient, and the average block temperature are
not available, a simpli ed analysis can be performed assuming that the ratio
of the miscellaneous heat losses
(
Q
miscell
aneous
miscellaneousmiscell
)
between two similar engines
is proportional to (V
E1
/V
E2
)
2/3
, where V
E1
and V
E2
are the displacements of
the two engines. In fact, V
E
2/3
is proportional to the characteristic heat transfer
area associated with the miscellaneous heat losses. The similarity between
the two engines refers to their respective con gurations such as inline or
Vee, EGR piping, and characteristic temperature in heat transfer losses.
The percentage of the miscellaneous heat losses can be approximated by
G
KT
V
JN
KT
V
h
KT
h
KT
th
h
KT
h
KT
1
11
KT
11
KT
1
2/
3
11
JN
11
JN
1
11
KT
11
KT
1
µ
Ê
Ë
Ê
Ë
Ê
Ê
Á
Ê
Ë
Á
Ë
Ê
Ë
Ê
Á
Ê
Ë
Ê
ˆ
¯
ˆ
¯
ˆ
ˆ
˜
ˆ
¯
˜
¯
ˆ
¯
ˆ
˜
ˆ
¯
ˆ
µ
ch
KT
ch
KT
11ch11
KT
11
KT
ch
KT
11
KT
E
V
E
V
EE
JN
EE
JN
11EE11
JN
11
JN
EE
JN
11
JN
ch
KT
ch
KT
11ch11
KT
11
KT
ch
KT
11
KT
E
V
E
V
h
2/
22/2
3
1
11
1
1
11
1/
3
=
h
v
h
v
th
BM
EP
BMEPBM
ht
11ht11
h
ht
h
11
h
11ht11
h
11
h
hthht
BM
EP
BMEPBM
VN
11
VN
11
K
VN
11
VN
11
1/
VN
1/
3
VN
3
EE
11EE11
VN
EE
VN
11
VN
11EE11
VN
11
htchht
11ht11ch11ht11
11E11
11
VN
11E11
VN
11
T
ht
T
ht
11ht11
T
11ht11
htchht
T
htchht
11ht11ch11ht11
T
11ht11ch11ht11
EEE
1
12.9
where the subscript 1 represents engine 1,
KT
V
h
KT
h
KT
11
KT
11
KT
2/
3
E1
V
E1
V
is an approximate
estimate of the miscellaneous heat losses, and J
E1
N
E1
/h
th1
is the fuel energy
rate. The h
th
is the engine brake thermal ef ciency. Similarly, the percentage
of miscellaneous heat losses of engine 2 can be approximated by
G
KT
VN
ht
KT
ht
KT
h
hthht
BM
EP
BMEPBM
2
22
ht22ht
KT
ht
KT
22
KT
ht
KT
h22h
hthht22hthht
22
VN
22
VN
1/
VN
1/
VN
3
VN
3
VN
2
µ
htchht
KT
ht
KT
ch
KT
ht
KT
ht22htchht22ht
KT
ht
KT
22
KT
ht
KT
ch
KT
ht
KT
22
KT
ht
KT
ht2ht
ht22ht2ht22ht
EE
VN
EE
VN
22EE22
VN
22
VN
EE
VN
22
VN
h
ht
h
ht
ht22ht
h
ht22ht
ht22ht2ht22ht
h
ht22ht2ht22ht
v
12.10
Diesel-Xin-12.indd 836 5/5/11 12:02:53 PM
837Diesel engine heat rejection and cooling
© Woodhead Publishing Limited, 2011
Therefore, the ratio of the percentages between engine 1 and engine 2 is
G
G
K
K
T
T
h
h
th
BM
EP
BMEPBM
BM
1
2
1
2
1
2
1
2
2
=
·
·
·
ch
T
ch
T
ch
T
ch
T
th
h
h
v
v
EP
EEPE
BMEBMEPBMEBM
N
N
V
V
1
2
1
2
1
1/
3
·
·
E
E
E
V
E
V
E
V
E
V
Ê
Ë
Ê
Ë
Ê
Ê
Á
Ê
Ë
Á
Ë
Ê
Ë
Ê
Á
Ê
Ë
Ê
ˆ
¯
ˆ
¯
ˆ
ˆ
˜
ˆ
¯
˜
¯
ˆ
¯
ˆ
˜
ˆ
¯
ˆ
12.11
When the two engines possess similarity, each of the rst several terms at the
right-hand side of equation 12.11 becomes 1. The following approximation
is then obtained as a scaling rule:
G
G
V
V
1
2
2
1
1/
3
=
E
V
E
V
E
V
E
V
Ê
Ë
Ê
Ë
Ê
Ê
Á
Ê
Ë
Á
Ë
Ê
Ë
Ê
Á
Ê
Ë
Ê
ˆ
¯
ˆ
¯
ˆ
ˆ
˜
ˆ
¯
˜
¯
ˆ
¯
ˆ
˜
ˆ
¯
ˆ
12.12
12.3 Characteristics of base engine coolant heat
rejection
12.3.1 Defi nition of base engine coolant heat rejection
Heavy-duty direct injection diesel engines usually have an oil cooler that
uses the engine coolant as cooling medium. Consequently, the engine coolant
heat rejection includes the heat removed by the oil cooler. Oil cooler heat
rejection consists mainly of two parts: the piston cooling and the engine
rubbing friction power loss. For system simulation, the total rubbing friction
can be obtained by subtracting the calculated pumping loss and accessory
power from the measured engine motoring power. The engine accessories
here include the water pump, the oil pump, the fuel pump, the alternator, and
the air compressor for the service brakes. A practical method for estimating
the motoring power is the Willan’s line (Heywood, 1988), where a plot of
fuel ow rate vs. BMEP is used to extrapolate to zero fuel ow rate in order
to determine the motoring mean effective pressure (MEP) or power.
A more detailed global friction model (Taraza et al., 2000) of individual
friction components (e.g., piston assembly, bearings, valvetrain) and auxiliaries
is desirable in engine system simulation to enhance accuracy. The model is
constructed with relatively simple physics governing the friction. The most
important design parameters of the components are included in the global
model. However, a more detailed friction model is needed for subsystem or
component design analysis. A very complex tribological model of piston-
assembly lubrication dynamics is shown in Chapter 10 as an example of an
advanced analytical design tool used in mechanical subsystem design.
Exhaust port heat rejection is a part of the base engine coolant heat rejection.
Its calculation is presented by Hires and Pochmara (1976), Rush (1976), and
Malchow et al. (1979). Cylinder head cooling studies are reported by Norris
et al. (1993). Piston cooling and heat rejection is reported by French (1972),
Pimenta and Filho (1993), Mian (1997), Varghese et al. (2005), and Agarwal
Diesel-Xin-12.indd 837 5/5/11 12:02:53 PM
838 Diesel engine system design
© Woodhead Publishing Limited, 2011
and Goyal (2007). Oil cooler design and testing information is provided by
SAE J1244 (2008), J1468 (2006), J2414 (2005), Adams (1999) and Adams
et al. (1999). Engine coolant heat rejection is discussed in Alkidas (1993),
Imabeppu et al. (1993), Shayler et al. (1996) Franco and Martorano (1999),
Campbell et al. (2000), and Luján et al. (2003). Heavy-duty diesel engine
lubrication system design is introduced by Kluck et al. (1986).
Coolant heat rejection can be specied in the following three forms: (1)
a heat transfer rate in kW or Btu/min (Fig. 12.6); (2) brake specic heat
rejection in, for example, kW/hp; and (3) a percentage of the fuel energy
that contributes to the coolant heat rejection. The percentage is usually
more suitable to compare different engines as a competitive benchmarking
parameter.
12.3.2 Low-heat-rejection engines
Coolant heat rejection is dominated by the large gas-to-wall thermal resistance
in the thermal boundary layer of the combustion chamber, rather than by the
smaller wall-to-coolant thermal resistance on the coolant side. Despite the
fact that ceramics have a much lower thermal conductivity than metals (by
one or two orders of magnitude), ‘adiabatic’ (or so-called low heat rejection
or LHR) engines using ceramic materials as the cylinder wall insulation do
not reduce heat rejection by an order of magnitude. The reason is that the
thermal resistance of the gas side plays a major role in the overall thermal
0 10 20 30 40 50 60 70 80 90 100 110 120 130 140 150
Fueling rate (lb/hr)
A range of coolant
heat rejection of
different diesel
engines
9000
8000
7000
6000
5000
4000
3000
2000
1000
Coolant heat rejection (Btu/min)
12.6 Competitive engine data of base engine coolant heat rejection
as a function of fuel flow rate.
Diesel-Xin-12.indd 838 5/5/11 12:02:54 PM
839Diesel engine heat rejection and cooling
© Woodhead Publishing Limited, 2011
resistance. Improving engine thermal efciency through in-cylinder heat
rejection reduction has been the key objective of LHR engines. It was found
that there was a reduction in the ignition delay and the premixed combustion
and an increase in the diffusion combustion duration in the LHR engines
when compared to the conventionally cooled engines. An adverse aspect of
LHR engines is that the much hotter cylinder wall temperature heats up the
induction air and results in a reduction in volumetric efciency and an increase
in NO
x
emissions. Overall, the ndings and conclusions on LHR engines
remain inconsistent and inconclusive. An improvement in LHR engines’
fuel consumption has been reported in the range of 4–10% (Jaichandar and
Tamilporai, 2003). The in-cylinder heat transfer characteristics of LHR
engines and their tribological impact are very complex and still not fully
understood by researchers.
Low heat rejection diesel engines have been extensively researched by
Morel et al. (1986), Assanis (1989), Kobori et al. (1992), Sun et al. (1994),
Yonushonis (1997). Kamo et al. (1987, 1996, 2000a, 2000b, 2003), Kamo
(2000), Sharma and Gaur (1990), Reddy et al. (1990), Woods et al. (1992),
Kimura et al. (1992), Schwarz et al. (1993), Bryzik et al. (1993), Jaichandar
and Tamilporai (2003, 2004), Tamilporai et al. (2003), Hergart et al. (2005),
Sutor et al. (2005), Saad et al. (2007), Rakopoulos and Giakoumis (2007),
and Yasar (2008).
12.3.3 Sensitivity of base engine coolant heat rejection
The most critical parameter for engine cooling system design is heat rejection
because it affects engine outlet coolant temperature, durability, and vehicle
front-end cooling packaging. Once the heat rejection is known, the cooling
medium ow rates of the radiator and the charge air cooler, the water pump
power, and the cooling fan power can be calculated relatively easily.
In addition to being characterized by the brake specic heat rejection, the
base engine coolant heat rejection can be characterized as a ‘percentage’
of fuel energy rate. The base engine heat rejection ‘percentage’ is affected
by cylinder heat transfer area, instantaneous heat transfer coefcient, and
engine friction. The relevant design or operating parameters include the
following: (1) cylinder bore diameter and stroke; (2) volume-to-surface ratio
of the combustion chamber, and engine compression ratio; (3) cylinder liner,
piston, and exhaust port design, especially the metal surface area exposed
to heat transfer; (4) swirl ratio and in-cylinder turbulence; (5) the ratio of
in-cylinder charge mass (air plus EGR) to fuel mass; (6) intake manifold gas
temperature; (7) fuel injection timing; (8) engine speed and load; (9) mean
piston speed; (10) water pump and oil pump power.
Note that the larger the in-cylinder charge-to-fuel ratio, the lower the base
engine heat rejection percentage. Moreover, the more retarded fuel injection
Diesel-Xin-12.indd 839 5/5/11 12:02:54 PM
840 Diesel engine system design
© Woodhead Publishing Limited, 2011
timing, the lower the percentage but the higher the BSFC (Fig. 12.1). The
coolant temperature, the coolant-side convection heat transfer coefcient,
and the cylinder head material (cast iron vs. aluminum alloy) have relatively
small impacts on coolant heat rejection. Good engine performance and low
heat rejection design require a small percentage of fuel energy lost to the
base engine coolant heat rejection. Figure 12.7 was obtained by engine
cycle simulations and it shows that the base engine heat rejection percentage
increases when the engine load decreases. Figure 12.8 shows that directionally
in-cylinder swirl reduction results in lower gas-side heat transfer coefcient
and hence lower base engine coolant heat rejection and lower BSFC. Figure
12.9 shows that the exhaust port length has an important impact on the port
heat rejection. Figure 12.10 shows the sensitivity of the base engine coolant
heat rejection with respect to engine power, EGR rate, air–fuel ratio, and
turbocharger efciency, .
12.4 Cooling system design calculations
12.4.1 Cooling system design considerations
Cooling system performance is important for engine performance. The
purpose of cooling is to maintain the engine component metal temperature
and the temperature gradient at appropriate levels (e.g., about 380°C for cast
iron and a lower temperature for aluminum alloys, 180–200°C for the top
piston-ring groove to prevent the deterioration of the lubricating oil lm).
Undercooled engines may suffer from the following problems: loss of material
Running with lower
air–fuel ratio and EGR rate
Rated power
Engine brake power
0
Base engine coolant heat rejection percentage in fuel energy
at constant engine speed
45%
40%
35%
30%
25%
20%
15%
Percentage of fuel energy lost to base
engine coolant heat rejection (excluding
EGR cooler heat rejection and 2.5% fuel
energy of miscellaneous heat losses)
12.7 Characteristics of base engine coolant heat rejection.
Diesel-Xin-12.indd 840 5/5/11 12:02:54 PM
841Diesel engine heat rejection and cooling
© Woodhead Publishing Limited, 2011
strength, high thermal strain, lubricant oil degradation, overheating and
scufng of the power cylinder components, heating the intake air, and lower
volumetric efciency. On the other hand, overcooled engines may exhibit
poor combustion and fuel economy, excessive heat rejection, high wear at
the piston rings and the liner, and high friction and noise. An excessively
low cylinder liner temperature, even below the condensation temperature of
the burnt gases, may cause liner corrosion. The maximum cooling capacity
of heavy-duty diesel engines is usually designed based on an in-vehicle
condition running at the rated power in summer at high altitude with air-
Note: In-cylinder low heat rejection design via turbulence
(swirl) reduction is the right direction for better BSFC and
heat rejection control.
16
15
14
13
12
11
10
9
8
Percentage of fuel energy lost to base
engine coolant heat rejection (%)
0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 1.1
Estimated in-cylinder turbulence level at TDC (related to swirl)
0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 1.1
Estimated in-cylinder turbulence level at TDC (related to swirl)
0.37
0.36
0.35
0.34
0.33
BSFC (lb/(hp.hr))
HD diesel engine at rated power with constant A/F ratio and peak
cylinder temperature
12.8 Impact of in-cylinder turbulence on base engine coolant heat
rejection.
Diesel-Xin-12.indd 841 5/5/11 12:02:54 PM