Available online at www.sciencedirect.com
Applied Thermal Engineering 28 (2008) 1395–1404
www.elsevier.com/locate/apthermeng
The influence of engine speed and load on the heat transfer
between gases and in-cylinder walls at fired and motored
conditions of an IDI diesel engine
Ali Sanli a, Ahmet N. Ozsezen
a
a,b
, Ibrahim Kilicaslan
a,*
,
Mustafa Canakci
a,b
Department of Mechanical Education, Kocaeli University, 41380 Izmit, Turkey
b
Alternative Fuels R&D Center, Kocaeli University, 41040 Izmit, Turkey
Received 4 July 2007; accepted 4 October 2007
Available online 16 October 2007
Abstract
In this study, the heat transfer characteristics between gases and in-cylinder walls at fired and motored conditions in a diesel engine
were investigated by using engine data obtained experimentally. For this investigation, a four-cylinder, indirect injection (IDI) diesel
engine was tested under different engine speeds and loads. The heat transfer coefficient was calculated by using Woschni expression correlated for the IDI diesel engines, and also using Annand and Hohenberg expressions. The temperature of in-cylinder gases were determined from a basic model based on the first law of thermodynamics after measuring in-cylinder pressure experimentally. The results
show that the heat transfer characteristics of the IDI diesel engine strongly depend on the engine speed and load as a function of crank
angle at fired and motored conditions.
2007 Elsevier Ltd. All rights reserved.
Keywords: Heat transfer coefficient; Heat flux; IDI diesel engine
1. Introduction
Heat transfer through the cylinder side walls is an
important process in determining overall performance, size
and cooling capacity of an internal combustion engine
(ICE). It affects the indicated efficiency because it reduces
the cylinder temperature and pressure, and thereby
decreasing the work transferred on the piston per cycle.
The heat loss (transfer) through the walls is in the range
of 10–15% of the total fuel energy supplied to the engine
during one working cycle [1].
From the in-cylinder point of view, the heat transfer
changes the local and instantaneous temperatures that
are of a critical importance in controlling NOx emission
formation. High temperatures lead to thermal stresses of
*
Corresponding author. Tel.: +90 2623032287; fax: +90 262 3032203.
E-mail addresses: ibrkilicaslan@hotmail.com, ikaslan@kocaeli.edu.tr
(I. Kilicaslan).
1359-4311/$ - see front matter 2007 Elsevier Ltd. All rights reserved.
doi:10.1016/j.applthermaleng.2007.10.005
material, and impact on fatigue failure limits of various
engine components, thus causing fatigue cracking or
deforming cylinder bore dimension and valve stems. Kept
below certain limits of the combustion chamber side wall
temperature is necessary to prevent the lubricating oil film
from deterioration and its viscosity from diminishing. In
order to avoid from pre-ignition and knock risks resulting
from overheated spark-plug electrodes and exhaust valves
in spark-ignition (SI) engines, the spark-plug and valves
must be kept cool. In addition to these, heat transfer to
the incoming air reduces volumetric efficiency [2].
There are a number of studies pertaining to the effects
on heat transfer characteristics of the various engine operating parameters. Alkidas [3] experimentally compared the
variations of the heat flux by using thermocouples mounted
at different locations of combustion side of cylinder head of
a single-cylinder SI engine at the motored and fired conditions. He showed that the magnitude of peak heat fluxes in
the combustion chamber increased with the engine speeds
1396
A. Sanli et al. / Applied Thermal Engineering 28 (2008) 1395–1404
Nomenclature
heat transfer surface area (m2)
constant pressure specific heat of cylinder gasses
(J/kg K)
D
cylinder bore (m)
h
instantaneous heat transfer coefficient (W/m2 K)
H/C
hydrogen/carbon
HCCI homogeneous charge compression-ignition
I/E
intake/exhaust
k
thermal conductivity of the gas (W/m K)
l
connection rod length (m)
m
mass of cylinder contents (kg)
Nan
swirl anemometer speed (rad/min)
Nu
Nusselt number
P
instantaneous in-cylinder pressure (kPa)
Pin
inlet pressure, 100 kPa
Pr
Prandtl number
q
heat transfer per unit area (MW/m2)
Re
Reynolds number
Ru
universal gas constant, 8.314 kJ/kmol K
S
cylinder stroke (m)
Tg
instantaneous in-cylinder gas temperature (K)
Tin
inlet temperature, 298 K
Tw
mean surface temperature of combustion chamber wall (K)
Tc
mean surface temperature of wall of coolant side
(K)
Up
(2Sn/60) mean piston speed (m/s)
Vs
swept volume (m3)
Vc
combustion chamber volume (m3)
vu
peripheral gas velocity (m/s)
A
Cp
at both conditions. He also compared the heat flux values
obtained by Woschni’s equation with the experimentally
measured ones and observed that the ones obtained by that
equation were in good agreement with the experimental
results. Watts et al. [4] showed that, at constant load for
SI engines, the average heat transfer through the cylinder
walls as a percentage of total fuel energy diminished almost
linearly with increasing engine speed first and thereafter did
not change at higher engine speeds. By using Woschni and
Annand equations, Karamangil et al. [5] investigated parametrically how the convective heat transfer coefficients for
gas side and coolant side varied with different engine operating parameters, such as the engine speed, compression
ratio, excess air coefficient, combustion duration, inlet
pressure and temperature. Rakopoulos et al. [6,7] carried
out the heat transfer analysis at some engines under various
engine operating parameters. In [6], it was examined the
heat release rates and the heat flux variations for both main
chamber and pre-chamber of a turbocharged IDI diesel
engine at the different load ranges. In [7], they studied the
w
x
effective gas velocity (m/s)
piston travel (m)
Greek symbols
D
difference
l
dynamic viscosity of cylinder gases (kg/ms)
q
density of cylinder gases (kg/m3)
/
equivalence ratio (relative fuel/air ratio)
c
specific heats ratio
k
relative air/fuel ratio
h
crank angle
Subscripts
c
coolant
g
gas
m
mean
p
piston
R
reference
w
wall
ABDC after bottom dead center
ATDC after top dead center
BBDC before bottom dead center
BTDC before top dead center
EVO exhaust valve opening
EVC exhaust valve closure
IVO
intake valve opening
IVC
intake valve closure
TDC top dead center
deg
degree of crank angle
rpm
revolutions per minute
instantaneous in-cylinder and exhaust pipe heat fluxes
under different speed and load ranges of a diesel engine
with air-cooled and one-cylinder. Shiling et al. [8] compared the measured instantaneous heat transfer coefficients
with the calculated values and found that the calculated
results were in a good agreement with the values obtained
experimentally.
As can be seen in the relevant literature, few heat transfer studies for IDI engines seem to have been performed
recently. Present work, intended for a contribution to the
open literature, deals with how to vary the heat transfer
coefficients and heat fluxes with the crank angle at different
speed and load ranges for fired and motored operations by
using the Woschni expression.
2. Heat transfer calculation between gases and in-cylinder
walls
Heat transfer phenomenon in the ICEs is a highly comprehensive and complex issue because of being transient,
A. Sanli et al. / Applied Thermal Engineering 28 (2008) 1395–1404
three dimensional, and the effect of periodic pressure and
temperature fluctuations of the charging. Furthermore,
when radiation heat transfer as well as convection heat
transfer in cylinder is considered, the issue becomes more
complicated. The heat transfer by radiation, taking place
from the high temperature solid soot particulates during
combustion process, has a dominant effect in diesel engines
and ranges about from 10% to 40% of total heat loss to the
walls, whereas this ratio is generally negligible and just
about 5–10% in SI engines [1,2,9].
Fig. 1 illustrates both the convective and radioactive
components, and the heat transfer between gas and in-cylinder wall, and conduction through the combustion chamber wall, and convection to the coolant. Mean gas and
coolant temperatures, Tmg and Tmc, are depicted with dotted lines [2].
Applying the assumption of steady heat transfer for the
present study, the heat flux can be calculated from the
Newton’s cooling law by multiplying the simultaneous heat
transfer coefficient with the temperature difference, i.e.
qðhÞ ¼ QðhÞ=A ¼ hðhÞðT g ðhÞ T w Þ
ð1Þ
Since the last 45 years, many models with global validity in
relation to the computing of the in-cylinder heat transfer
coefficient have been proposed by a number of investigators, e.g. Woschni [10], Annand [11] and Hohenberg [12]
and so on, whose models have relied on dimensional analysis for turbulent flow that correlates the Nusselt, Reynolds
and Prandtl numbers, which is
Nu ¼ aReb Prn ;
ð2Þ
where Pr has been omitted because of changing little in
gases, so its effect can be included in the ‘‘a coefficient’’.
In particular, the Woschni correlation has frequently
been used in the heat transfer studies with proper constants
in today’s SI and diesel engines, and also it has been correlated conveniently even for HCCI engines recently [13].
Tg
Conduction
T mg
Convection
Convection+Radiation
T mc
Tw
Tc
Instantaneous heat transfer coefficient adopted from
Woschni is calculated by
hðhÞ ¼ 3:26P ðhÞb T g ðhÞ0:751:62b Db1 wðhÞb
Cylinder
wall
Coolant side
Fig. 1. Schematic of the overall heat transfer process in the cylinder.
ð3Þ
He assumed the ‘‘b exponent’’ is 0.8 and emphasized that
effective gas velocity, w, consists of two contributions.
The first contribution is scaled with mean piston motion
and swirl, and the second contribution is related to turbulence effects and DP pressure rise resulted from combustion, and of course this contribution includes the
influence of radiation [1,14]. The term
wðhÞ ¼ ð2:28 þ 0:308vu =U p ÞU p þ C
T RV s
DP
P RV R
ð4Þ
for the compression and expansion strokes
wðhÞ ¼ ð6:18 þ 0:417vu =U p ÞU p
ð5Þ
for scavenging period, namely the induction and exhaust
strokes. Where vu is defined as
vu ¼
pDN an
60
ð6Þ
and Nan can be calculated from 0.7B below the cylinder
head in a steady flow test [15,16], Up is mean piston speed,
Vs is stroke volume, TR, PR and VR are respectively taken
as the temperature and pressure of working gases and total
volume on piston at a chosen reference state, either at intake valve closing (IVC) or at the start of combustion, here
IVC. DP, P(h) Pm(h) is instantaneous difference in pressure between the firing and motoring engine at the same
crank angle. In the present study, the motoring pressure
Pm was measured experimentally instead of a calculation
method suggested by Watson and Janota [15] and was observed the different results than that of the model at the two
different engine speeds as indicated in Fig. 2. At the start of
combustion, C coefficient has been correlated by Woschni
as 3.24 · 103 for both the diesel engines and SI engines,
whereas as 6.22 · 103 for IDI diesel engines because of
higher turbulence effects.
The other approach for the instantaneous heat transfer
coefficient, developed by Annand, can be expressed as
hðhÞ ¼ a
k b
Re
D
ð7Þ
The value of ‘‘a constant’’, which represents the level of
convective heat transfer, varies with intensity of charging
motion and combustion chamber design. At normal combustion, it varies from 0.25 to 0.8 and increases directly
with increasing the intensity of charging motion. The index
b varies from 0.7 to 0.8 [17]. k thermal conductivity of the
gas as a function of Tg can be calculated from following
formula [2,18];
kðhÞ ¼ 3:17 104 T g ðhÞ
Gas side
1397
0:772
ð8Þ
or from
kðhÞ ¼
C p lðhÞ
Pr
ð9Þ
1398
A. Sanli et al. / Applied Thermal Engineering 28 (2008) 1395–1404
7000
6500
6000
model in Ref [15]
5500
1500 rpm
5000
1000 rpm
Pressure, kPa
4500
Motored
condition
4000
3500
3000
2500
2000
1500
1000
500
0
-180 -165 -150 -135 -120 -105 -90 -75
-60 -45
-30 -15
0
15
30
45
60
75
90
105 120 135 150 165 180
Crank Angle, deg
Fig. 2. Comparison of the pressures from motored operation and Watson–Janota model.
Eq. (9) for k in the present work was used. Cp is constant
pressure specific heat of cylinder gases. l, viscosity of combustion products, is a function of temperature and equivalence ratio, and almost independent of the pressure [2]. It is
computed from following equation:
0:7
lðhÞ ¼ 3:3x107 T g ðhÞ =ð1 þ 0:027/Þ
ð11Þ
For the specific heats ratio c, one generally takes it between
1.3 and 1.35 values. Here, c was found from / equivalence
ratio (k = /1) and this variable has been continuously updated during the different engine operating conditions.
c ¼ 1:4 0:16/
ð12Þ
Reynolds number is formed with a characteristic speed
equal to the Up mean piston velocity, a characteristic length
equal to the D cylinder bore, q gas density and l dynamic
viscosity as follows:
qðhÞU p D
ReðhÞ ¼
lðhÞ
ð13Þ
The density of cylinder contents during time between IVC
and EVO is mainly a function of instantaneous cylinder
volume
m
q¼
ð14Þ
V ðhÞ
Assuming perfect gas behavior,
m¼
PVM mix
Ru T g
V ðhÞ ¼ V c þ
ð10Þ
Pr, Prandtl dimensionless number, is nearly constant for
ordinary gases and equal to 0.74 [19]. However, we have
calculated it from following equation due to being measured / 6 1
Pr ¼ 0:05 þ 4:2ðc 1Þ 6:7ðc 1Þ2
Mmix is the molar mass of the mixture of air and fuel. The
instantaneous cylinder volume and area as a function of
crank angle are given by
ð15Þ
AðhÞ ¼
pD2
xðhÞ
4
pD2
4
qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
pDS
R þ 1 cosðhÞ þ R2 sin2 ðhÞ
þ
2
ð16Þ
ð17Þ
x(h) represents for distance from TDC, and thus according
to the crank angle
qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
xðhÞ ¼ l þ R R cosðhÞ þ ðl2 R2 sin2 ðhÞÞ
ð18Þ
Hohenberg made some modifications in the Woschni’s formula, such as the instantaneous cylinder volume instead of
bore, the effective gas velocity, and the exponent of the
instantaneous gas temperature. Hohenberg’s equation
therefore is described as
hðhÞ ¼ 3:26P ðhÞ0:8 T g ðhÞ0:4 V ðhÞ0:06 ðU p þ cÞ0:8 ;
ð19Þ
where c is the calibration constant that Hohenberg suggested to be 1.4 for the engine he studied [12,20].
As the cylinder gas temperature, the gases far away from
the wall are used. Hence the instantaneous gas temperature
was calculated by assuming the first law of thermodynamics and using the cylinder pressure;
c1
P ðhÞ c
T g ðhÞ ¼ T in
:
ð20Þ
P in
c can be computed as a function of equivalence ratio by
means of Eq. (12) as cited already for hydrocarbon fuels.
A. Sanli et al. / Applied Thermal Engineering 28 (2008) 1395–1404
Tw the mean combustion chamber wall temperature
depends on the engine speed, load, equivalence ratio, start
of combustion, charge motion, inlet temperature, wall
material, and the coolant and combustion temperatures.
It involves apparently too sophisticated to predict the wall
temperature and is generally chosen a constant value during all operating conditions, such as 350 K [21], 650 K
[22]. However, in this study, as presented by Cheung and
Heywood [23], Tw was determined depending on the /
equivalence ratio obtained experimentally in order to reach
better agreement, instead of a random value.
For / < 0:833;
T w ¼ 400 K;
1399
Table 1
Technical specifications of 1.8 VD BMC IDI diesel engine
Engine type
Water-cooled, four strokes and naturally
aspirated
Number of cylinder
Bore/stroke
Connecting rod
length
Compression ratio
IVO/IVC
EVO/EVC
I/E valve diameter
Injection pump
Static injection timing
Maximum power
4
80.26/88.9 mm
158 mm
21.47:1
8 deg BTDC/44 deg ABDC
50 deg BBDC/10 deg ATDC
36.55/30.78 mm
Mechanically controlled distributor type
18 deg BTDC
38.8 kW (52 HP) @ 4250 rpm
For 0:833 < / < 0:9; T w ¼ 425 K;
For 0:9 < /; T w ¼ 450 K:
hm the mean heat transfer coefficient can be computed via a
cycle simulation to the calculated instantaneous heat transfer coefficient and then integrated as following:
Z 720
1
hm ¼
hðhÞdðhÞ
ð21Þ
720 0
Likewise, the mean cylinder gas temperature
Z 720
1
T mg ¼
hðhÞT g ðhÞdðhÞ:
720hm 0
ð22Þ
Thus, mean heat flux is
qm ¼ hm ðT mg T w Þ:
ð23Þ
3. Apparatus and procedure
Engine tests were performed in a water-cooled, naturally
aspirated, four-stroke, IDI diesel engine. Engine specifications are given in Table 1. Engine torque was adjusted by
means of a Motosan water-cooled hydraulic dynamometer
and was read through a SP 200 load cell with 1 g sensitivity
and 0–200 kg ranges. The engine speed was measured by a
magnetic speed sensor referenced with respect to TDC
point. The start of combustion was determined from sharp
rising in curvature of the heat release rate that happens at
ignition instant [24,25].
The lambda k, relative air–fuel ratio or in other words
the inverse of the equivalence ratio /, was measured by
using an exhaust emission test equipment. A Kistler model
6061B water-cooled piezoelectric pressure transducer with
a Kistler 5015A 1000 charge amplifier located into No. 1
cylinder head was used for measuring the cylinder pressure.
A software program installed in a PII computer was used.
The data acquisition system collected the pressure data at
every 0.25 crank angle, and the average of consecutive 50
cycles for fired and 20 cycles for motored conditions were
obtained. The experimental study was carried out at the
Engine Test Laboratory of Department of Automotive
Technologies in Kocaeli University. The schematic experimental set-up is shown in Fig. 3.
Exhaust emission
test equipment
Computer
Data
acquisition
system
Pressure
signal
Speed
counter
ENGINE
Dynamometer
Load
cell
Load
panel
Fig. 3. The engine experimental set-up.
1400
A. Sanli et al. / Applied Thermal Engineering 28 (2008) 1395–1404
Fig. 4. The motored operation of the test engine.
The first stage concerning the fired conditions was performed under the different engine speed and load ranges.
The experiments were carried out at 20 Nm, 40 Nm and
full load ranges for 1000 and 1500 rpm engine speeds. Fuel
used was No. 2 diesel (H/C = 1.78) with 56.5 cetane number and 42930 kJ/kg lower heating value. The second stage
concerning the motored conditions, in order to account for
the heat transfer behaviors without fuel, was conducted at
the same test engine by forced via a 15 kW AC electrical
motor assembled to engine gearbox shaft. The speeds of
the electrical motor were adjusted with a variable speed
drive. The photograph of this assembly is seen in Fig. 4.
4. Results and discussion
4.1. Heat transfer coefficients
The histories of the heat transfer coefficients calculated
by using the Woschni, Annand, Hohenberg expressions
and others have been shown that all being different in the
studies of the present and literature [1,5,16,21]. Fig. 5
shows the histories of heat transfer coefficients calculated
by Woschni, Annand and Hohenberg expressions for
1000 rpm at full load (84.9 Nm).
Their maximum values vary from 1700 to 6900 W/m2 K.
It is believed that the reason of this variation has originated
from the ‘‘a coefficient’’ (0.49), the ‘‘C coefficient’’
(6.22 · 103) and ‘‘b exponent’’ indexes (0.8 for Woschni,
0.7 for Annand) used in the Woschni and Annand equations, and from the modifications in the Hohenberg equations, such as decreasing of the gas velocity and pressure
exponent value. As mentioned previously, because the
Woschni’s equation has especially been correlated for IDI
diesel engines, it has been thought to be more reasonable
utilizing this equation rather than the others in order to
show the obtained results in subsequent sections.
4.2. The heat transfer characteristics at the 20 Nm, 40 Nm
and full loads
Typical histories of heat transfer coefficients and heat
fluxes as a function of crank angle are shown in Figs. 6–8
for each engine speed and the load ranges, 20 Nm,
40 Nm and full load.
The highest heat transfer rates to the walls generally
take place during the compression and expansion periods
near TDC and are strongly influenced by the gas pressure
and temperature. During other periods, i.e. intake and
exhaust periods, the value of heat flux changes nearly to
a small negative [3,28]. The variations near TDC for the
all heat fluxes are comprehensibly presented and enlarged
as indicated in the Figs. 6–8. Comparisons of the results
obtained at different engine test conditions show that the
7000
6500
6000
Woschni
Heat transfer coefficient, W / m2K
5500
Hohenberg
5000
Annand
4500
1000 rpm
Full load
4000
3500
3000
2500
2000
1500
1000
500
0
-180 -165 -150 -135 -120 -105 -90 -75
-60 -45 -30 -15
0
15
30
45
60
75
90 105 120
135 150 165 180
Crank Angle, deg
Fig. 5. Comparison of Woschni, Annand and Hohenberg heat transfer coefficients.
A. Sanli et al. / Applied Thermal Engineering 28 (2008) 1395–1404
magnitude of peak heat fluxes varies relatively with the
engine speeds and loads. The highest difference among
these values was seen as high as 0.55 MW/m2 between
20 Nm and full loads at 1500 rpm. Lowest peak heat flux
value was 1.64 MW/m2 for 20 Nm at 1500 rpm. On the
other hand, the highest peak heat flux was at 1500 rpm full
load (88.8 Nm), which is 2.19 MW/m2 as shown in Figs. 6
1401
and 8. As a result, this variation caused to an increase of
33.5% on the peak heat flux. Similar trends were also seen
for heat transfer coefficients. The maximum h value was
seen as 7450 W/m2 K, at 7 deg ATDC at 1500 rpm full load
condition and the minimum h value was seen as 5135 W/
m2 K, at 6.5 deg ATDC at 20 Nm load of the same speed.
This increasing of h value obviously leads to the increase in
5500
4.0
3.8
5000
compression
expansion
exhaust
intake
3.6
3.4
3.2
3.0
4000
2.8
1.6
1000 rpm
3500
2.6
1500 rpm
1.2
2.4
0.8
2.2
3000
0.4
2.0
0.0
2500
-20
-10
0
10
20
30
1.8
40
1.6
load=20 Nm
2000
Φ =0.392 at 1000 rpm
Φ =0.401 at 1500 rpm
1500
ignition
1000
1000 rpm heat transfer coefficient
1500 rpm heat transfer coefficient
1000 rpm heat flux
1500 rpm heat flux
500
0
Heat flux, MW / m2
Heat transfer coefficient, W / m2K
4500
1.4
1.2
1.0
0.8
0.6
0.4
0.2
0.0
-180 -150 -120 -90 -60
-30
IVC
0
30
60
90
TDC
120 150 180 210 240 270 300 330 360 390 420 450 480 510 540
Crank Angle, deg
EVO
IVO EVC
Fig. 6. Variations of the heat transfer coefficients and heat fluxes at 20 Nm load condition.
4.0
6500
3.8
6000
compression
expansion
exhaust
intake
3.6
3.4
5500
3.0
4500
1000 rpm
1500 rpm
4000
2.0
2.8
1.6
2.6
1.2
2.4
0.8
3500
2.2
0.4
2.0
0.0
-20
3000
-10
0
10
20
30
40
1.8
2500
load=40 Nm
2000
Φ =0.452 at 1000 rpm
Φ =0.531 at 1500 rpm
1.6
1.4
1.2
1.0
1500
1000 rpm heat transfer coefficient
1500 rpm heat transfer coefficient
1000 rpm heat flux
1500 rpm heat flux
1000
500
0
-180 -150 -120 -90 -60 -30 0
30
IVC
TDC
60
90
0.8
0.6
0.4
0.2
0.0
120 150 180 210 240 270 300 330 360 390 420 450 480 510 540
EVO
IVO
EVC
Crank Angle, deg
Fig. 7. Comparison of heat transfer coefficients and heat fluxes at 40 Nm load condition.
Heat flux, MW / m2
Heat transfer coefficient, W /m2K
3.2
5000
1402
A. Sanli et al. / Applied Thermal Engineering 28 (2008) 1395–1404
4.0
8000
3.8
7500
compression
expansion
exhaust
intake
7000
3.4
6500
3.2
3.0
6000
2.4
5500
2.8
2.0
1000 rpm
1500 rpm
2.6
1.6
5000
4500
4000
1.2
2.4
0.8
2.2
0.4
2.0
0.0
-20
3500
-10
0
10
20
30
1.8
40
1.6
full load
Φ =0.884 at 1000 rpm
Φ =0.97 at 1500 rpm
3000
2500
2000
1000 rpm heat transfer coefficient
1500 rpm heat transfer coefficient
1000 rpm heat flux
1500 rpm heat flux
1500
1000
500
0
-180 -150 -120 -90 -60 -30
IVC
Heat flux, MW / m2
2
Heat transfer coefficient, W / m K
3.6
1.4
1.2
1.0
0.8
0.6
0.4
0.2
0.0
0
TDC
30
60
90
120 150 180 210 240 270 300 330 360 390 420 450 480 510 540
EVO
Crank Angle, deg
IVO
EVC
Fig. 8. Comparison of heat transfer coefficients and heat fluxes at full load condition.
heat flux. While the heat transfer coefficient slightly
decreased with an increase in the engine speed at low load,
it increased relatively with the engine speed at the higher
loads. The maximum h and heat flux values for 1000 rpm
are 5212 W/m2 K and 1.7 MW/m2, respectively, as seen
in Fig. 6.
Fig. 7 shows the heat transfer coefficient and heat flux
histories at 40 Nm constant load. At this load, as different
from that occurred at 20 Nm, the peak heat transfer coefficient slightly increased with an increase in the engine
speed, whereas the peak heat fluxes were nearly a constant
value. The peak heat flux values were higher compared
with that 20 Nm, and their magnitude were almost the
same as 2.10 MW/m2 ATDC for both 1000 and
1500 rpm. Similarly, the maximum h value occurred at
the levels of 6255 W/m2 K at 3.75 deg ATDC for
1000 rpm and 6345 W/m2 K at 7 deg ATDC for
1500 rpm. When the engine load was changed from
20 Nm to 40 Nm at 1000 rpm constant speed, the heat flux
and heat transfer coefficient increased by 23.5% and 20%,
respectively.
The peak heat transfer coefficient of the 1500 rpm was
relatively higher than that of 1000 rpm, at full load, as indicated in Fig. 8. Increase of engine speed from 1000 to
1500 rpm led to approximately the increasing of 7.6% for
the peak heat transfer coefficient, from 6927 W/m2 K to
7450 W/m2 K. On the other hand, for the heat fluxes with
increasing engine speed at this load, there were hardly ever
varying like that 40 Nm, 2.19 MW/m2 value for each speed
at full load. A possible explanation for this is partly the
increasing of the mean wall temperature owing to the
increasing of equivalence ratio and thus decreasing of
(Tg Tw) temperature difference. At all loads, the variations in the heat fluxes with increasing engine speed are
clearly attributed to the variations of heat transfer coefficient resulted from the changes in the gas pressure, temperature and velocity and turbulence severity. As pointed out
by Jafari and Hannani [26], increasing engine speed leads
to a longer combustion period in terms of crank angle,
causing an increase in the overlap of the burning time with
the expansion stroke. Thus, increasing engine speed causes
a decrease in the heat loss per cycle but increases the heat
loss per unit time.
Moreover, one of the mainly significant characteristics
from the noticed results is that the heat fluxes and heat
transfer coefficients exhibit the orderly sequences with the
increasing loads for each speed. These observations can
easily be seen if the histories on the figures presented above
simply are examined. Since the increasing of loads means
the richer mixtures, the increasing of charge energy leads
to the increase of the ones just mentioned. The engine
speed affects the heat transfer to the in-cylinder walls and
its augmentation enhances the characteristic velocity of
the flow, and consequently turbulence motion. These
obtained observations come to close results reported by
other researchers attempting on this subject pertaining to
the effect of engine operating parameters on the cylinder
heat transfer [3–7,14,16,26–30].
4.3. The heat transfer characteristics for motored (unfired)
operating
At motored conditions, the heat transfer magnitudes
decline considerably because of no combustion. The engine
850
800
compression
750
700
Heat transfer coefficient, W / m2K
650
600
550
500
450
400
350
300
250
200
150
100
50
0
-180 -150 -120 -90 -60
-30
0
30
0.50
0.48
0.46
expansion
exhaust
intake
0.44
0.42
0.40
0.38
0.36
0.34
0.32
0.30
0.28
1500 rpm heat transfer coefficient
0.26
1000 rpm heat transfer coefficient
0.24
1500 rpm heat flux
0.22
1000 rpm heat flux
0.20
0.18
Motored condition
0.16
γ = 1.4, Tw = 300K
0.14
0.12
0.10
0.08
0.06
0.04
0.02
0.00
60 90 120 150 180 210 240 270 300 330 360 390 420 450 480 510 540
1403
Heat flux, MW / m2
A. Sanli et al. / Applied Thermal Engineering 28 (2008) 1395–1404
Crank Angle, deg
Fig. 9. Comparison of the heat fluxes and heat transfer coefficients by using Woschni expression at the motored operation.
was operated by using the electrical motor shown in the
Fig. 4 at 1000 and 1500 rpm speeds without the fuel to
observe the difference of the combustion effect on the heat
transfer characteristics. For motored conditions, c was
chosen 1.4 due to assuming the ideal gas and Tw value
was chosen 300 K for each speed.
The variations of heat flux and heat transfer coefficient
with crank angle are shown in Fig. 9. Compared with those
of the fired operation of the engine, the values of motored
peak heat fluxes and heat transfer coefficients are seen to be
fairly less. They are almost 0.21 MW/m2 for 1000 rpm,
0.28 MW/m2 for 1500 rpm, and 592 W/m2 K for
1000 rpm, 818 W/m2 K for 1500 rpm. All peak values of
heat flux (HF) and heat transfer coefficient (HTC) obtained
in the tests are given in Table 2.
According to these results, the heat flux increasing at
20 Nm, 40 Nm and full load for 1000 rpm are respectively
8.1, 10 and 10.4 times higher, and for 1500 rpm are respectively 5.85, 7.5 and 7.82 times higher than those of unfired
conditions. Likewise, the heat transfer coefficient increasing at 20 Nm, 40 Nm and full load for 1000 rpm are respectively 8.8, 10.6 and 11.7 times higher, and for 1500 rpm are
6.27, 7.75 and 9.1 times higher than those of unfired
conditions.
Table 2
The peak values of HF and HTC obtained in the tests
Unfired
condition
20 Nm
40 Nm
Full
load
h (W/m2 K)
1000 rpm
1500 rpm
592
818
5212
5135
6255
6345
6927
7450
q (MW/m2)
1000 rpm
1500 rpm
0.21
0.28
1.7
1.64
2.1
2.1
2.19
2.19
Furthermore, as different from those at fired conditions,
the histories of heat transfer coefficients and heat fluxes
were seen rising the orderly sequences at motored condition, and all peak values also occurring at the same crank
angle, TDC, because of no combustion effect.
5. Conclusions
The objective of this study was to understand how the
heat transfer characteristics between the gases and the incylinder combustion chamber walls of an IDI diesel engine
would vary with respect to the engine speeds and loads at
fired and motored conditions. Based on the results calculated by using the Woschni expression, the following conclusions can be drawn.
The increasing of engine speed at low constant load
(20 Nm) has a little decreasing effect on both the in-cylinder
peak heat fluxes and heat transfer coefficients, whereas the
increasing of engine speed at higher loads (40 Nm and full
load) has somewhat increasing effect on the in-cylinder heat
transfer coefficients; however, the heat fluxes remain about
the same for the cases mentioned. Peak heat fluxes did
hardly ever change due to decreasing the difference between
the gas temperature and the mean wall temperature.
The increase in engine load at constant speed has a
major effect on the peak heat fluxes and heat transfer coefficients over the combustion chamber wall surfaces. The
heat flux and heat transfer coefficient histories exhibit
clearly that the differences between their peak values are
to lessen as the load increases at the constant speed.
At motored conditions, the peak heat fluxes and heat
transfer coefficients increased with the engine speed, rising
from 1000 rpm to 1500 rpm led to 33.3% increase of heat
1404
A. Sanli et al. / Applied Thermal Engineering 28 (2008) 1395–1404
transfer coefficient, to 38.18% increase of heat flux. Compared the peak heat flux values of fired with those of unfired,
increasing at 20 Nm, 40 Nm and full load for 1000 rpm are
respectively 8.1, 10 and 10.4 times higher, and as for
1500 rpm, they are respectively 5.85, 7.5 and 7.82 times
higher. Likewise, the heat transfer coefficient increasing at
20 Nm, 40 Nm and full load for 1000 rpm are respectively
8.8, 10.6 and 11.7 times higher, and as for 1500 rpm, they
are 6.27, 7.75 and 9.1 times higher than those of unfired
condition.
Acknowledgements
The experimental data used in this study were obtained
from the project supported by The Scientific & Technological Research Council of Turkey (TUBITAK), Project No.
104M372. The authors are grateful to the institutes and the
individuals at the engine test laboratory who were involved
in making this work possible.
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