A Four-Stroke Homogeneous Charge Compression Ignition Engine Simulation for Combustion and Performance Studies

17 Pages·0.354 MB·English

by  Fiveland S.B., Assanis D.N.

2000-01-0332
A Four-Stroke Homogeneous Charge Compression Ignition
Engine Simulation for Combustion and Performance Studies
Scott B. Fiveland and Dennis N. Assanis
W. E. Lay Automotive Laboratory
University of Michigan
Ann Arbor, MI 48108
Copyright © 2000 Society of Automotive Engineers, Inc.
ABSTRACT
A computer simulation of the Homogenous Charge (Homogeneous Charge Compression Ignition Diesel
Compression Ignition (HCCI) four-stroke engine has Combustion, Diesel Fumigation, MULDIC (Multiple
been developed for combustion and performance Staged Diesel Combustion, PREDIC (Premixed Direct-
studies. The simulation couples models for mass, Injection Combustion), and PCIC (Premixed
species, and energy within a zero-dimensional Compression Ignited Combustion). Engine types have
framework. The combustion process is described via a ranged from two-stroke to four stroke configurations with
user-defined chemical kinetic mechanism. The a variety of fuels such as diesel, gasoline, methanol,
CHEMKIN libraries have been used to formulate a stiff natural gas, and hydrogen. The HCCI process
chemical kinetic solver suitable for integration within a essentially involves a premixed fuel/air mixture that is
complete engine cycle simulation, featuring models of inducted into the cylinder at equivalence ratios that can
gas exchange, turbulence and wall heat transfer. For vary from lean to stoichiometric [6, 7]. Once within the
illustration, two chemical kinetics schemes describing cylinder, the homogeneous fuel/air charge is then
hydrogen and natural gas chemistry have been compressed until ignition commences. Ignition leads to a
implemented in the code. The hydrogen scheme is a very rapid combustion phase where all heat is released
reduced one, consisting of 11 species and 23 reactions. approximately in 10-35o.
The natural gas chemistry is described via the GRI- The premixed charge, compression ignition HCCI
mechanism 3.0 that considers 53 species and 325 engine concept promises to combine the advantages of
reactions, including NOx chemistry. Computations are both the Direct Injection, Compression Ignition (DICI)
first carried out in a variable volume bomb to engine and the premixed charge, Spark-Ignited (SI)
demonstrate variations in ignition with temperature, engine, while eliminating their drawbacks. One benefit
pressure, equivalence ratio, and composition. that can be gained over the current heterogeneous DICI
Subsequently, the complete cycle simulation is exercised engines is the elimination of the fuel rich zones that are
to demonstrate the variation in output parameters to directly responsible for pollutant formation, especially
charge inlet temperature and effective compression ratio. particulates [7]. In addition, the homogeneous lean burn
Overall, this study demonstrates the importance of operation will yield lower gas temperatures and hence
coupling detailed chemistry descriptions with physical NOx, as compared to both SI and DICI counterparts.
models of the HCCI engine processes. Furthermore, unthrottled part load operation eliminates
pumping losses leading to improved fuel economy over
INTRODUCTION the SI engine [1]. In addition, compression ignition
eliminates SI knock associated with autoignition of fuel-
Homogeneous Charge Compression Ignition
air mixture in the end zone, ahead of the advancing
(HCCI) is currently under widespread investigation due
flame front. As the HCCI concept operates on the
to its potential to lower NOx and particulate emissions
premise of autoignition, this allows the use of elevated
while maintaining high thermal efficiency [1, 2, 3, 4, 5].
compression ratios (approximately 20-25:1), unlike in SI
Throughout the years it has endured many names in the
engines. The combination of lean burning, which is
literature: ATAC (Active Thermo-Atmosphere
thermodynamically attractive and reduces heat transfer
Combustion), LHC (Lean Homogeneous Combustion),
losses, and high compression and thus expansion ratios
CIHC (Compression Ignited Homogeneous Charge
contributes to indicated thermal efficiencies that can
Combustion), AR (Active Radical Combustion), HCDC
approach 55%. Residual gas fraction is also reduced
with HCCI operation, thus improving volumetric integrating flexible chemical kinetic libraries with models
efficiency, combustion and performance [5]. Overall, of turbulence-based heat transfer and gas exchange
HCCI promises to deliver both high thermal efficiencies processes for a four-stroke cycle. The paper is arranged
and reduced emissions. as follows. The HCCI model formulation will be
Although recent investigations into HCCI presented first. This will include a brief development of
combustion appear promising [4, 5, 6, 7, 8, 9, 10, 11, 12] the governing equations and various engine submodels
several problems with the HCCI combustion concept (i.e. gas exchange, combustion, heat transfer) as they
reappear throughout the literature. Most stem from the apply to a system with well understood chemistry. The
fact that the HCCI concept gives up two combustion sensitivity of the combustion submodel to changes in
control aspects. First, the timing of ignition is not pressure, temperature, equivalence ratio, and natural
controlled, neither indirectly by fuel injection as in a DICI gas composition will be demonstrated first in a
engine, nor directly by the spark as in an SI engine. fundamental, adiabatic reactor. The behavior of the each
Second, the rate of heat release is also not controlled of the submodels will then be detailed over the engine
neither by the rate of fuel injection as in a DICI engine, cycle, and contrasted with other treatments of cyclic
nor by finite turbulent flame propagation as in an SI processes reported in the literature. Subsequently,
engine. As a result, the near constant volume parametric studies of ignition and engine performance for
combustion event leads to a very rapid rate of heat varying manifold temperature, pressure, and geometric
release, thus promoting high mechanical stresses [8]. In compression are performed for both hydrogen and
particular, controlling the ignition event at higher loads is natural gas using the full cycle simulation.
a widely noted problem.
In an effort to understand how mixture preparation MODELING ASSUMPTIONS
and in-cylinder thermodynamic conditions affect the
The compression ignition engine simulation of
chemical kinetics, models of varying resolution have
Assanis and Heywood [15] has provided a solid zero-
been developed [4, 7, 9, 10, 11, 12]. These modeling
dimensional framework for formulation and
efforts provide a good basis for exploring the HCCI
implementation of the governing equations for the HCCI
combustion phenomena. However, in published zero-
model. The simulation is currently written in a single
dimensional simulations, detailed chemistry has typically
cylinder version, primarily because fundamental studies
been coupled to models of turbulence, heat transfer, and
lend themselves to this configuration. Thermodynamic
gas exchange that are not predictive over the full engine
properties are assumed uniform throughout the chamber
cycle. Heat transfer models have been based on data
volume. The engine simulation is a sequence of four-
not relevant to HCCI operation, and in many cases the
stroke processes. The gas exchange process is
gas exchange process has been neglected by specifying
governed by quasi-steady, one-dimensional flow
conditions at a specific point in the engine cycle. For
equations that are used to predict flow past valves. The
instance, Smith et al. [7], Van Blarigan and Goldborough
compression event is defined from Intake Valve Closing
[5], as well as Kusaka et. al. [12] used the Woshni heat
(IVC) to a transition point prescribed when chemical
transfer correlation [13] to predict heat transfer. The
reactions become important. The combustion event for
Woschni model [13] was developed from a regression of
the HCCI simulation differs from those of the SI and DICI
direct injection diesel engine data and is essentially a
types. As a result of its premixed nature and
linearized radiative/convective relation that will greatly
compression ignition principle, the rate of combustion is
over-predict heat transfer in a lean burn, premixed, non-
strictly limited by the chemical kinetics. In the reaction
sooting engine. Furthermore, studies by Poulos and
regime, the combustion will follow that of finite rate
Heywood [14] as well as Assanis and Heywood [15]
kinetics and its heat release will be governed by detailed
have shown that, since turbulence intensity has a large
chemistry. Hence, the combustion event has been
cyclic variation, heat transfer correlations based on
modeled using the Chemkin libraries [16], adapted for a
multiples of mean piston speed do not provide a truly
variable volume plenum and accounting for heat transfer
predictive capability. In addition, the computational work
effects. The evolution of heat release and species is
of Smith et. al. [7] was only carried out over the closed
governed by a user-defined kinetic scheme. Later in
part of the cycle, starting with bottom dead center of the
expansion, the reacting flow mixture will attain chemical
intake stroke, with no gas exchange. Studies of closed
equilibrium in response to the changing in-cylinder
cycle processes are useful for evaluating chemistry, but
conditions, and eventually the composition will become
do not capture intake and exhaust jet flows.
frozen. At this point the mixture is again non-reacting
Consequently, velocity and length scales of in-cylinder
and the chemical energy source term tends to zero.
flow structures are not accurately represented; thus, wall
heat transfer losses and pre-ignition boundary conditions
CONSERVATION EQUATIONS
can be inaccurate.
The objective of this work is to develop a full cycle To define a thermodynamic state within the cylinder,
simulation model of the HCCI engine that would integrate two independent properties and the mixture composition
complex chemistry with physical models of the in- need to be known. Consequently, our formulation will
cylinder processes. Emphasis will be placed on track the evolution of mass, species, temperature, and
pressure throughout the engine cycle. The general Expanding Eq. (3), and applying the continuity equation
equations governing mass, species, and energy will be yields,
dFeigvuerleo p1e. d I t fiosr aas svuamrieadb lteh avt othluem weo rrkeiancgt oflru, ida sb eshhaovwesn ains Y(cid:6)i =(cid:229) (cid:231)(cid:231)Ł(cid:230) m(cid:6)mj ł(cid:246)(cid:247)(cid:247) (Yij- Yicyl)+W iWr mw (4)
an ideal gas. The heat transfer is governed by turbulent j
pipe flow, but derives its characteristic velocity as a
function of mean piston speed, mean velocity, and which is the final form of the species conservation
turbulent intensity, which is found from a zero- equation.
dimensional k-e formulation.
CONSERVATION OF ENERGY – The generalized
energy equation for an open thermodynamic system may
be written as:
d(mu) dV dQ
= - p + ht +(cid:229) m(cid:6) h (5)
j j
dt dt dt
(cid:9)(cid:10)(cid:11) (cid:9)(cid:10)(cid:11) (cid:9)(cid:10)(cid:11) (cid:9)(cid:12)j (cid:10)(cid:12)(cid:11)
Internal Displacement Heat Enthalpy
Energy Work Transfer Flux
Rewriting the first law equation in terms of rate of change
of enthalpy yields,
mh(cid:6) =(cid:229) m(cid:6) h - Q(cid:6) +P(cid:6)V- m(cid:6)h (6)
j j w
j

At this point in the development, an expression needs to
(cid:6)
be developed for h that relates it to the change in
Figure 1 Energy Transfers Associated with Cylinder
mixture temperature. Assuming a single phase, multi-
Control Volume.
component mixture of ideal gases, its enthalpy, h, is
defined as
CONSERVATION OF MASS – The rate of change
of mass within any open system is the net flux of mass h =(cid:229) Yjhj (7)
across the system boundaries. j
dm =(cid:229) m(cid:6) (1) and h =h(T,P,Yi) (8)
j
dt
j
where the subcript ‘j’ refers to each of the component
species present in the mixture. Taking partial differentials
CONSERVATION OF SPECIES – Equations
with respect to pressure, temperature and composition:
tracking the evolution of species within the combustion
chamber will be developed on a mass basis
corresponding to the definition in Eq. (2), dh = ¶ h dT + ¶ h dp +(cid:229) ¶ h dYi
dt ¶ T dt ¶ p dt ¶ Y dt
P,Yi T,Yi i i P,T,Yi„ j
m
Y = i (2) (9)
i
m
where the partial change of enthalpy with respect to
where ‘m’ denotes the total mass within the control
pressure at constant temperature and composition is
cylinder. The species equations are deduced from their
zero, but pressure effects are accounted for in
multi-dimensional counterparts by neglecting species
determining changes in composition with allowance for
diffusion terms, consistent with the zero-dimensional
dissociation. Hence, Eq. (9) becomes
assumption.
d(mY) (cid:6) (cid:6) dh = (cid:229) Y C dT +(cid:229) h dYj = C dT +(cid:229) h dYj
dt i =(cid:229) mjYij+Sgen dt j j pj dt i j dt p dt i j dt
j
(3) (10)
(cid:6)
S =W W v
gen i mw To relate the pressure gradient to the temperature
change, the equation of state is used in its differential
form. By manipulating the thermodynamic equation of lift and discharge coefficients can be specified or
state for an ideal gas, we get: predicted. The values for cylinder pressure are updated
by solving the system of state differential equations
R(cid:6) V(cid:6) P(cid:6) m(cid:6) T(cid:6) through the cycle [18].
= + - - (11)
R V P m T
Ø 1 ø
An expression can now be inserted for the rate of Œ 1 1(cid:236) Ø g - 1ø (cid:252) 2œ
cthhearnmgoed yonfa mthice degpaesn dceonncset,a in.et .:R by considering its m(cid:6) i,e = ŒŒŒ CdAeff (cid:231)(cid:231)Ł(cid:230) RuPTo2oW (cid:247)(cid:247)ł(cid:246) 2(cid:231)(cid:231)Ł(cid:230) PPTo (cid:247)(cid:247)ł(cid:246) g (cid:239)(cid:239)(cid:237) g 2-g 1ŒŒŒ 1- (cid:231)(cid:231)Ł(cid:230) PPTo ł(cid:246)(cid:247)(cid:247) g œœœ (cid:239)(cid:239)(cid:253) œœœ
Œ (cid:238) º ß (cid:254) œ
( ) º ß
R =(cid:229) Y R (12)
j j
j (18)
Taking partial differentials with respect to pressure, COMBUSTION – The combustion process in a
temperature and composition, homogenous charge compression ignition engine
exhibits no fundamental mixing or entrainment, which
(cid:230) (cid:246) normally control the combustion event for direct-injection
dR =(cid:229) (cid:231)(cid:231) Y R(cid:6) (cid:229)+ ¶ R Y(cid:6) (cid:247)(cid:247) (13) as well as spark ignition engines. As a result the rate of
dt j (cid:231)Ł j j j ¶ Yj P,T,Y j(cid:247)ł heat release is solely driven by the chemical kinetic
i„ j reaction rates. To appropriately model the combustion
event, it was necessary to describe the evolution of heat
The first term on the R.H.S. of Eq. (13) can be
release via a suitable chemical kinetic mechanism, and a
recognized as the partial expansion of the gas constant
well-matched, stiff chemical kinetic solver. CHEMKIN
for a pure substance. Under the problem assumptions,
was selected due to the fact that it is a widely accepted
this term can be shown to be identically equal to zero,
and used kinetic solver by many researchers in a range
which is consistent with references [16] and [17].
of combustion studies [10, 19]. Simulation development
has been done in Fortran 90 within the Visual
R(cid:6) =(cid:229) R Y(cid:6) (14) Environment on a PC 333 Mhz Pentium II. A source
j j
j program for a variable volume reactor was written to
drive the CHEMKIN libraries. The driver model includes
Using Eqs. (11) and (14), we can obtain an equation that the effects of heat transfer, as well as options for
relates the rate of change of pressure to the rate of different engine configurations. The program can also
change of temperature, i.e.: operate in an adiabatic standalone mode for fundamental
studies.
Ø (cid:229) (cid:6) ø The rates of creation/destruction of chemical
R Y
P(cid:6) =PŒŒ i i i +m(cid:6) +T(cid:6) - V(cid:6) œœ (15) species are modeled using mass-action kinetics, where
the specific reaction rate constants exhibit a strong
Œ R m T Vœ
º ß temperature dependence. An elementary reaction that
involves ‘K’ chemical species in ‘I’ reactions can be
represented in the form,
Susbstituing Eq. (15) into Eq. (6) yields an equation for
the rate of change of temperature for an ideal gas,
K K
reacting mixture: (cid:229) u ' c (cid:219) (cid:229) u '' c (19)
ki k ki k
k=1 k=1
T(cid:6) = A1ŒŒŒºØ - (cid:231)(cid:231)(cid:231)Ł(cid:230) (cid:229) i hi- vP(cid:229)Ri Ri(cid:247)(cid:247)(cid:247)ł(cid:246) Y(cid:6)i- mm(cid:6) B+m1(cid:231)(cid:231)Ł(cid:230) (cid:229) j m(cid:6)jhj+PV(cid:231)(cid:231)Ł(cid:230) - VV(cid:6) (cid:247)(cid:247)ł(cid:246) +Q(cid:6)w(cid:247)(cid:247)ł(cid:246) œœœßø W k =(cid:229) I (u ''ki- u 'ki)qi (20)
i=1
(16)
where A = Cp - PTv and B=h- Pv (17) q =k (cid:213)K [X ]u 'ki - k (cid:213)K [X ]u 'k'i (21)
i f k r k
MODELING OF RECIPROCATOR PROCESSES ik=1 ik=1
GAS EXCHANGE – The one-dimensional quasi-
The specific reaction rate constant, ‘k’ follow the
steady flow model is used to model flow through both the
Arrhenius dependence where,
intake and exhaust valves during the gas exchange
processes. Equation (18) is a function of discharge
coefficient, valve area, gas properties, and pressure kf =AiTb i exp(cid:231)(cid:230) - Ei (cid:247)(cid:246) (22)
differential across each orifice. The values for both valve i Ł RT ł
THERMODYNAMIC PROPERTY TREATMENT -- dK 1 m(cid:6)
The fluid in the cylinder is constantly undergoing a = m(cid:6) iVi2 - P, - K e (27)
change in mixture composition. Once the composition is ,dt (cid:9)2(cid:12)(cid:10)(cid:12)(cid:11) Production (cid:9)(cid:10)m(cid:11)
Change Mean TKE Exhausted
determined, the partial mixture properties can be MKE MKE MKE
summed and appropriately weighted in accordance with
their mass or mole fractions. This method very simply dk m(cid:6)
allows thermodynamic property calculation for mixtures = P, - m,e - k e + A, (28)
containing residual gas, exhaust gas recirculation, ,dt Production Dissipation (cid:9)(cid:10)m(cid:11) Source
unburned gaseous fuel, air etc. The thermodynamic Change TKE TKE Exhausted TKE
TKE TKE
property treatment will employ the NASA curve-fits for
specific heat, enthalpy, and entropy:
(cid:230) ( )3 (cid:246)
(cid:231) (cid:247)
Copk = (cid:229)5 a Tn- 1 (23) e = u"'3 =(cid:231)(cid:231) 2k3"m 2 (cid:247)(cid:247) (29)
nk (cid:231) (cid:247)
R n=1 Ł ł
where the constant of integration an+1,kRis the The turbulence production is modeled based on
enthalpy of formation at 0 K. turbulence production in the boundary layer flow over a
flat plate [15, 20]:
RhTokk =a1k +n(cid:229)5=1ann,kTkn- 1 + aT6kk (24) P=m tCb (cid:231)Ł(cid:230) ULł(cid:246)(cid:247) 2 ; m t = Cmm ke2 (30)

sok =a lnT + (cid:229)4 an+1,kTn +a (25) where Cm =0.09(Universal) and Cb is an adjustable
1k k k 7k constant.
R n=1 n

Large scale eddies can be represented by a geometric
HEAT TRANSFER MODEL -- The heat transfer
length scale defined as
model that is utilized was described by Assanis and
Heywood [15]. This model is a subset of the k-e model
V
under the isotropic turbulence assumption. It uses an "= (31)
energy cascade between the mean kinetic energy and A
b
the turbulent kinetic energy. The eddy dissipation is
modeled via dimensional argument. This model accounts
Rewriting Eq. (30) yields
for the effects of mean piston speed, mean charge

motion, and turbulence intensity on the heat transfer
1
coefficient. P= 2Cm Cb K(cid:231)(cid:230) k (cid:247)(cid:246) 2 (32)
The definitions for both mean and turbulent kinetic 3 " Ł mł
energy are given by Eq. (26). (cid:231)(cid:230) 2(cid:247)(cid:246) 2
Ł 3ł
1
K = mU2
2 Rapid distortion theory (RDT) is employed to account for
(26)
k = 3mu'2 rapid changes in density. This amplification source term
2 was added to the transport equation for turbulent kinetic
energy
The equations governing the cascade model are shown
in Eqs. (27)-(29). It is noted that these equations take on du'
A =3mu' (33)
the classical form for transport phenomena, having an
dt
unsteady term, a convective term, and a source/sink
term. The mean flow kinetic energy equation describes
Since RDT assumes negligible dissipation of the
the history of the initial kinetic energy. Over time, this
turbulent kinetic energy, conservation of mass and
mean kinetic energy is dissipated to large scale
angular momentum of each eddy is assumed [15, 20].
turbulence, whose evolution is described by the turbulent
kinetic energy equation. Finally the eddy dissipation,
1 1
which occurs on a molecular level, dissipates the " (cid:230) r (cid:246) 3 u' (cid:230) r (cid:246) 3
=(cid:231)(cid:231) o (cid:247)(cid:247) : =(cid:231)(cid:231) (cid:247)(cid:247) (34)
turbulent kinetic energy. It is modeled via dimensional "o Ł r ł uo' Ł r o ł
considerations.
du' u' dr DVODE, is then used to deal with the kinetic equations
dt = 3r dt (35) that involve a wide range of reaction time scales, until the
species concentrations are no longer changing. A
flowchart showing the structure of the engine simulation
2 r(cid:6)
A = k (36) is shown in Figure 2.
3 r After initializing the state variables, the program
proceeds through each engine process. The main
The rate of change of density is computed at each program controls the solution process. During each
timestep. From the above definitions, the characteristic timestep, the main program calls the pertinent integrator,
velocity is computed as which then calls one of the four main process routines.
Each of these process routines determines what
1 equations are important and integrate them over each
Vch =ŒŒØ U2 +u'2 +(cid:231)(cid:231)Ł(cid:230) V2p (cid:247)(cid:247)ł(cid:246) 2œœø 2 (37) ttihmee FsOteRp.T TRhAeN c 7o7m/9p0u tceor mmpoudteerl hlaansg ubaegeen. implemented in
º ß
BEHAVIOR OF SIMULATION SUBMODELS
The heat transfer coefficient can be computed from
correlations of Nusselt number as a function of Reynolds COMBUSTION - The combustion process in the
number. The interested reader can refer to Assanis and HCCI engine is driven by chemical kinetics. For
Heywood [15] for details. demonstration purposes both hydrogen and natural gas
were used as the fuels within the full single cycle
METHOD OF SOLUTION simulation. Many researchers consider hydrogen as the
ultimate fuel of the future [22]. It has been under study
The mathematical model results in a set of ordinary for hybrid applications due to its high energy content and
differential equations. These equations must be low emissions potential. In addition, hydrogen exhibits
integrated simultaneously. Due to the incorporation of unique fuel characteristics, which include high flame
chemical kinetics, two different integrators are used to speed, ignitibility of lean mixtures, and high effective
ensure computational efficiency. When the chemical octane [23]. Furthermore, its well understood kinetic
reactions are frozen, the set of equations is integrated scheme, which involves only a moderate number of
using the standardized, non-stiff, predictor-corrector reactions, makes it a good candidate for modeling
technique, ODERT [21], based on a modified form of the studies that can be used for demonstration purposes.
Adams Pece method. When pre-combustion reactions Likewise, natural gas, which contains a composition on
begin to occur late in the compression stroke, the the order of 85-90% methane, 2-10% ethane, and 1-4%
program switches integrators. A "stiff" ODE integrator, propane has received attention in recent years due to its
Start Read Input
Initialize
Integrate Intake 1. Engine Geometry
Main Program 2. Valve Data
Integrate Compression
IVO --- IVC 3. Flow Rates
4. Heat Transfer
Integrate Combustion
IVC --- Ignition 5. Hydro-Carbon Formation
Integrate Expansion 6. Engine Performance
Ignition -- EVO 7. Thermodynamic Properties
Integrate Exhaust 8. Transport Properties
EVO -- IVO
Write Output
Y N
N Y
Combustion
Converge
Finite Rate Specified
Kinetics ROHR
End
Figure 2. Logic Structure of the Thermo-kinetic HCCI Engine Simulation
"clean burn" potential. Its chemical kinetic mechanism is
0.022
reasonably well characterized making it a good
Correlation Coefficient (R) = 1.
candidate for fundamental studies that can lead to a c) 0.021
e
working engine concept. To demonstrate the flexibility of S
thhyed rsoimgeunla ticohne mtoi satcryc epwt aasl tedrneastcivrieb ekdin ewticit hs chae mreedsu, ctheed p/dt ( 0.02
d
scheme available in CHEMKIN that considers 11 species n
o
and 25 reactions [16]. For natural gas, the GRI- ed 0.019
s
Mechanism 3.0 was input into CHEMKIN; the detailed a
B
mechanism considers upwards of 53 species and 325 y 0.018
a
reactions [24]. Del
n 0.017
o
Combustion in a Variable Volume Plenum - cti
u
d
Before using the mechanism in the full simulation, it was n 0.016
I
studied in a fundamental variable volume, adiabatic
reactor to verify that it reproduced reported ignition 0.015
trends for variations in inlet temperature, inlet pressure, 0.015 0.016 0.017 0.018 0.019 0.02 0.021 0.022
and the addition of higher order hydrocarbons. Correctly Induction Delay based on OH or Ho (Sec)
2
quantifying the onset of ignition is critical to developing a
predictive simulation since it has a large effect on engine Figure 3 Correlation between induction delays based on
output. For our study, the time of occurrence of the peak maximum pressure gradient and peak OH or HO
2
concentration of either OH or HO was used as the concentrations for methane combustion.
2
ignition criterion. This is consistent with Glassman [25]
who notes that that the presence of H, O, OH, and HO
2
abstract H radicals from methane and hence promote 1.2 108
chain propagation. Note however, that other studies 400 K
define the onset of ignition based on the occurrence of 450 K
1 108
the maximum pressure gradient [25]. Even though 500 K
mixture conditions throughout the compression process
change until the onset of ignition occurs, a useful Pa) 8 107 600 K
quantitative indicator of induction delay (sec) can be e ( 800 K
ur
defined as the time that lapses between Bottom Dead ss 6 107
e
Center (BDC, start of compression) and the onset of Pr
ignition. Figure 3 examines the correlation between
4 107
induction delays based on maximum pressure gradient
and peak OH or HO concentrations for methane
combustion in a variable 2volume combustor over a range 2 107
of inlet temperatures and pressures. The compression
ratio is 15 and the speed is 1500 rpm. Remarkably, the 0
correlation coefficient between the two induction delays 0.012 0.014 0.016 0.018 0.02 0.022 0.024 0.026
is 1, suggesting that both ignition criteria are consistent Time (Sec)
for the case of HCCI combustion.
Figure 4 Effect of initial mixture temperature on ignition
Having defined the ignition criterion, induction delay
and pressure profile in a variable volume chamber.
studies were carried out as the initial conditions of the
Methane combustion with equivalence ratio of 0.5 and
mixture in the variable volume combustor were varied.
initial mixture pressure of 1.5 bar.
The first case study varied the initial (BDC) temperature
of a methane-air mixture from 400 K to 800 K. The
higher inlet pressures generating more free radicals that
equivalence ratio was maintained at 0.5, and the initial
in turn accelerate the chemical reactions.
mixture pressure at BDC was 1.5 bar. The resulting
Composition of natural gas can vary widely from
pressure traces are shown in Fig. 4. As anticipated, the
state to state, as well as slightly from cycle to cycle [26].
induction delay decreased with increasing temperature
As a result many fundamental studies have been carried
(see Fig. 5). It was further noted that methane did not
out to understand how ignition behaves under varying
ignite for an intake manifold temperature of 400 K,
composition [9, 10, 27, 28]. Compositions were chosen in
consistent with previous studies [eg., 10].
a range consistent with Westbrook and Pitz [27]. A
Next the initial pressure at BDC was varied from 1
methane-air mixture at 1.5 bar and a temperature of 500
bar to 3 bar, while keeping the temperature fixed at 500
K was used as the baseline case. Ethane and then
K, the compression ratio fixed at 15, and the mixture
finally propane was gradually added to the mixture in
composition as above. As shown in Fig. 6, ignition time
increased quantities. Computed induction delays as a
is advanced with higher inlet boost. This is a result of the
function of composition are shown in Fig. 7. Consistent
increased trapped air and fuel concentrations at these
0.022 0.0193
Baseline: 1.0 CH
4
0.021
0.0192
c) 0.02 c)
e e0.0191
S S
ay (0.019 ay ( 0.97 CH4 + 0.03 C2H6
el el
D D 0.019
n n 0.95 CH + 0.05 C H
o0.018 o 4 2 6
cti cti
u u
d d0.0189
In0.017 In 0.90 CH + 0.10 C H
4 2 6
0.0188
0.016
0.80 CH + 0.10 C H +0.1 C H
4 2 6 3 8
0.015 0.0187
0.0012 0.0014 0.0016 0.0018 0.002 0.0022 0.0024 0 1 2 3 4 5 6
(Temperature)-1 1/K Test Case
Figure 5 Correlation of induction delay as a function of Figure 7 Variation in induction delay under different
initial mixture temperature. Methane combustion in a compositions of methane, ethane and propane. Natural
variable volume chamber with equivalence ratio of 0.5 gas combustion in a variable volume chamber with initial
and initial mixture pressure of 1.5 bar. mixture conditions of 1.5 bar and 500 K, and an
equivalence ratio of 0.5.
Combustion in Engine Simulation -- The
2.5 108
behavior of the combustion submodel was further
1 BAR explored within a full engine cycle simulation. Primary
engine specifications are summarized in Table 1. Unless
2 108 1.5 BAR otherwise noted, for the engine combustion study
2 BAR reported in this section, as well as for subsequent
Pa) studies illustrating the behavior of other submodels: (i)
e (1.5 108 3 BAR hydrogen was used as the fuel, (ii) the engine geometric
sur compression ratio was fixed at 15, (iii) intake manifold
s
e pressure was 1.5 bar, intake manifold temperature was
Pr 1 108 425 K, mixture equivalence ratio was 0.3. and (iv) engine
speed was 1500 rpm.
5 107 Table 1 Engine Specifications
Engine Type 4-Stroke
Displacement 12.7 Liter
0 Bore 13cm
0.01 0.015 0.02 0.025 0.03 Stroke 16cm
Time (Sec) Con. Rod Length 26.93cm
Intake Valve Opening 685 deg *
Intake Valve Closing 220 deg *
Figure 6 Variation of onset of ignition with changing
Exhaust Valve Opening 485 deg *
initial pressure of mixture. Methane combustion in a
Exhaust Valve Closing 795 deg *
variable volume chamber with equivalence ratio of 0.5
* Relative to TDC Intake
and initial mixture temperature of 500 K.
A pressure-volume diagram, over the full engine
cycle, is shown in Fig. 8a. Clearly, the combustion is very
with Westbrook and Pitz [27], Fraser et. al. [28], Naber
close to constant volume. This rapid rate of heat release
et. al. [9], and Agarwal and Assanis [10], increased
is in agreement with the work of Van Blarigan and
concentrations of ethane and propane reduced induction
Goldsborough [5]. The pressure and temperature history
delays, which is a result of increased radical pools
is shown in Fig. 8b. As will be shown in subsequent
promoting chain branching reactions.
sections, the rapid variation of ignition as well as rapid
1000 0.0015 0.0006
OH H2O2
(a)
0.0005
100 ctions (-) 0.001 HOO2 NNO 0.0004Minor S
ssure (bar)10 es Mole Fra0.0005 HprOec2u, r sHo, ra fnodr iOgnHit,i opnroduction 0.0003pecies Mole
LN Pre 1 Major Speci 0 00..00000012 Fractions (-)
0
NO Production
-0.0005 -0.0001
00.1.0001 LN Vo0l.u0m01e (m3) 0.01 0.008 0.0085 0.009 0.0095Time0.01 (Sec)0.0105 0.011 0.0115 0.012
120 2500
Figure 9 Variation of major and minor species during
(b) hydrogen combustion in an engine operating with an inlet
100
2000 manifold temperature of 425 K, an inlet pressure of 1.5
bar, and an equivalence ratio of 0.3.
80
ar) Ignition Point 1500 Tem
b
sure ( 60 Temperature peratu The first is due to the rapid blowdown event that occurs
es 1000 re when the exhaust valve opens, and the second is
Pr (K
40 Pressure ) consistent with the piston motion. Figure 10b shows the
calculated mean flux velocities for both the intake and
500 exhaust velocities. The large flow velocities have a
20
direct effect on the in-cylinder flow field and resulting
heat transfer coefficient. Note that negative velocities
0 0 correspond to reverse flows due to an adverse pressure
200 250 300 350 400 450 500 gradient across the valve.
Crank Angle (deg) Including the gas exchange processes within the
Figure 8 Cyclic characteristics of a four-cycle engine cycle simulation, as opposed to specifying
running on hydrogen at an inlet manifold temperature of thermodynamic conditions at BDC, allows the simulation
425K, an inlet pressure of 1.5 bar, and an equivalence to be used for variable valve timing studies. For
ratio of 0.3. (a) pressure-volume diagram; (b) variation demonstration purposes, the exhaust valve closing time
of gas pressure and temperature during combustion. was varied from 725o to 795o ca-deg to demonstrate the
effect that it has on ignition and the resulting engine
heat release greatly effect engine performance. The performance. The test case was run using hydrogen
species formed during hydrogen combustion are tracked fuel, an inlet pressure of 1.5 bar, and an inlet manifold
as a function of time, and are shown in Fig. 9. Peaks in temperature of 500K, although the results are
H, OH, and HO can be discerned near the ignition point. qualitatively the same at other operating conditions.
2
Note that NO emissions are near zero, a result of lean Figure 11 shows results for ignition timing, indicated
combustion. thermal efficiency, and volumetric efficiency. Closing the
exhaust valves early results in a reduction in scavenging
GAS EXCHANGE -- Profiles detailing the rate of mass efficiency. This causes more internal EGR to reside in
flow through the engine intake/exhaust valves, computed the cylinder after the exhaust valves close, and leads to
using the one-dimensional, quasi-steady, compressible higher mixture temperatures after induction. The
flow model, are shown in Fig. 10a. It is noted that the elevated temperatures and active radicals in the internal
intake valve mass flow rate follows the path of the piston residual gas cause earlier ignition timing. Since ignition
motion, in agreement with Assanis and Heywood [15]. timing is advanced into the compression stroke, the
On the other hand, the exhaust flow exhibits two peaks thermal efficiency is greatly reduced with earlier ignition.
1.2 TURBULENT FLOW MODEL-- The eddy length
(a) scale of the turbulence is governed as a function of the
m 1 distance between cylinder head and piston, as noted in
wmaximu0.8 Intake Exhaust Einq F. i(g3u3r)e. 1It2sa v. aMrioarteio nd etthariolsu gcha nth bee e fnoguinnde icny cAlses iasn sish oawnnd
o
Fl Heywood [15]. Figure 12b shows the variation in the
s 0.6 Exhaust Blowdown mean flow velocity, turbulent intensity, and the overall
s
a
M characteristic velocity. It is noted during the intake stroke
w/ 0.4 that the mean kinetic energy is initially large and then
o
Fl decreases in response to piston motion. It decreases
s 0.2
s more slowly during the compression stroke. The mean
a
M
flow then increases during the exhaust blowdown, a
0
result of the large pressure potential across the valve.
The large scale turbulence levels are shown to be very
-0.2
high during the intake flow. Later in the intake event,
-100 0 100 200 300 400 500 600 700
Crank Angle (deg) turbulence decays since the rate of dissipation, which is
governed by a scaling argument, due to viscous shear
400
stresses is larger than the rate of turbulence production.
Exhaust Valve Velocity
(b) The use of RDT to account for rapid density changes
300 results in a slight amplification of the turbulence intensity
during combustion. The characteristic velocity is a
m/s)200 Intake Valve Velocity scaling argument made up of mean piston speed, mean
y ( gas velocity, turbulent intensity, and during the exhaust,
cit a blowdown velocity. The variation of both velocity flow
o100
el scales as well as the piston speed, has a direct effect on
V
ve the characteristic velocity as well as the subsequent heat
Val 0 transfer process.
-100 HEAT TRANSFER MODEL -- Once the
characteristic velocity is found, the film coefficient for
Reverse Flow To Intake Manifold
heat loss can be calculated. The heat transfer coefficient
-200
-100 0 100 200 300 400 500 600 700 and heat transfer per unit area are shown in Figure 13. It
Crank Angle (deg) is noted that it experiences a strong variation over the
Figure 10 (a) Mass flows and (b) velocities through the cycle. This is due to a combination of variation in gas
intake and exhaust valves for hydrogen-fueled engine properties (i.e. thermal conductivity and viscosity) as well
operating at 1500 rpm. as the characteristic velocity and the macroscale of
turbulence. The second peak in the film coefficient is due
%) 120 360 to the large exhaust blowdown.
y ( Volumetric Efficiency A heat transfer study was performed to examine
c
n
e the effect of different heat transfer models on ignition.
ci 100 355
Effi Specifically, the Woschni heat transfer model [13], that
ncy, Volumetric 6800 Ignition Point 334550 Ignition Point (d hwtasrtadaausinsad s bybfc.aee otTerim cnh m pecua ooscrmdeoeedmdbl u ipntsboa tty rihoi tshenAoe mncs assza saejwoenroer iwsi-rt ydea i asmorn ufue dnpsn ueasHbdito le ianasy haswc elo aodmk o b-Hdpea rCseb[eC1sal5sisIn i]esoe. dtn ui ndrhT aietehhtasioee,t
e e
Effici g) ooff 1475,0 aKn, ianlltehto purgehs ssuimreil aorf t3r ebnadrs, afonldlo awn a itn loetth teerm oppeerraattuinreg
mal 40 340 conditions. The pressure and temperature profiles are
er
h Thermal Efficiency shown in Figs. 14a and 14b. Clearly, the heat transfer
T
d. model has a pronounced effect on the ignition point. As
In 20 335
anticipated, the adiabatic case ignites first because the
720 730 740 750 760 770 780 790 800
lack of heat transfer promotes a higher mean cylinder
Exhaust Valve Closing (deg)
gas temperatures and hence earlier ignition. But what is
Figure 11 Variation of volumetric efficiency, thermal most surprising is that the Woschni model predicts the
efficiency, and ignition timing for different exhaust valve ignition point 8 degrees later than that by Assanis and
closing times. Hydrogen-fuelled engine operating with Heywood [15]. This discrepancy is attributed to the fact
intake manifold pressure of 1.5 bar and intake manifold that the Woschni model was developed for DICI diesel
temperature of 500 K. engines by fitting data to the correlation form. In other