International journal of automotive technology, tập 10, số 5, 2009

Effects of CO 2 injection parameters injection timing, quantity, pressure on HCCI combustion and emission characteristics were investigated.. Either advancing CO 2 injection timing or in

International Journal of Automotive Technology , Vol 10, No 5, pp 529 − 535 (2009)

529

EFFECT OF DIRECT IN-CYLINDER CO2 INJECTION ON HCCI COMBUSTION AND EMISSION CHARACTERISTICS

S QU, K DENG * , L SHI and Y CUI

Key Laboratory for Power Machinery and Engineering of Ministry of Education, Shanghai Jiao Tong University,

Shanghai 200240, China

(Received 31 December 2007; Revised 24 October 2008)

ABSTRACT− Fuel injection during negative valve overlap period was used to realize diesel homogeneous charge compression ignition (HCCI) combustion In order to control the combustion, CO 2 in-cylinder injection was used to simulate external EGR Effects of CO 2 injection parameters (injection timing, quantity, pressure) on HCCI combustion and emission characteristics were investigated Experimental results revealed that CO 2 in-cylinder injection can control the start of combustion and effectively reduce NO x emission Either advancing CO 2 injection timing or increasing CO 2 injection quantity can reduce peak cylinder pressure and mean gas temperature, delay the starts of low temperature reaction (LTR) and high temperature reaction (HTR), and lower pressure rise rate; NO x emission was reduced, while smoke, HC, and CO emissions increased Since the combustion phase was improved, the indicated thermal efficiency was also improved Injection pressure determines the amount of disturbance introduced into the cylinder Generally, with the same injection quantity, higher injection pressure results in higher momentum flux and total momentum Larger momentum flux and momentum has a stronger disturbance to air-fuel mixture, resulting in a more homogeneous mixture; therefore, larger injection pressure leads to lower

NO x and smoke emissions

KEY WORDS : Homogeneous charge compression ignition, Gas injection system, Injection timing, Injection quantity

1 INTRODUCTION

Diesel engines are widely used in trucks and buses due to

their superior fuel economy, high torque at low speed,

durability, and reliability, and are increasingly used as car

engines However, conventional diesel combustion has the

problem of high NOx and particulate matters (PM)

emi-ssions, and a trade-off relationship exists between NOx and

PM emissions According to a recent simulation study

(Kitamura et al., 2002) on n-heptane (a simulation

sub-stitute for diesel), NOx is formed in the high temperature

region (T > 2200 K) and oxygen rich zones, while PM is

formed in a certain temperature range (1500 K~2500 K)

and fuel rich zones (φ> 2) A technique that can reduce one

type of emissions would increase the other type of

emissions For example, EGR can effectively reduce NOx

emission, but it would increase PM emission

With increasingly stringent emission regulations, many

techniques have been used to reduce diesel engine

emi-ssion, one of which was to use air jet to promote air-fuel

mixing during the diffusion combustion phase to reduce

soot emission Kamimoto et al. (1983) added an air cell in

the cylinder head of a direct injection (DI) diesel engine

He found that air was pushed into the air cell during the

compression stroke and was injected into the main chamber

after the cylinder pressure reached its peak The air jetstirred the stagnant flame and promoted soot oxidation.The results showed that the air-cell system effectivelyreduced soot emission at medium and high loads Nagano

et al. (1991) added a plunger and a spring in the air cell.When the air was pushed into the air cell, it also pushed theplunger, and the spring was compressed so that themomentum flux of the air increased when air was injectedinto the main chamber Several studies were made by Kurtz

et al. and Choi et al. (Choi and Foster, 1995; Kurtz et al.,

1998, 2000) by using auxiliary gas injection to increase cylinder mixing during the latter portion of combustion toreduce soot emission Effects of gas injection direction,compositions, momentum, and injection tim-zing on reduc-ing soot emission were studied

in-New combustion concepts have also been developed,such as HCCI, which can simultaneously reduce NOx andsmoke emission HCCI combustion is intensively studieddue to its emission reduction potential for DI diesel engines.Unlike conventional diesel combustion, air and fuel ofHCCI combustion are premixed, and the homogeneousmixture is auto-ignited at multiple points throughout thecylinder NOx emission is dramatically reduced due to thelow combustion temperature resulted from a lean air-fuelmixture, and PM emission is reduced due to a well pre-mixed mixture and absence of fuel rich zones However,HCCI combustion can only be realized in low and medium

*Corresponding author. e-mail: kydeng@sjtu.edu.cn

engine operating regimes because it is hard to control the

combustion rate at high load

In studies in the last few years, internal and external

EGR were widely used to control HCCI combustion The

former is mainly used in gasoline-fueled HCCI engines to

dilute the homogeneous mixture in order to obtain

suffici-ently high temperatures for auto-ignition (Onishi et al.,

1979) The latter is mainly used to control diesel-fueled

HCCI combustion (Peng et al., 2003; Okude et al., 2004;

Hardy et al., 2006; Kimura et al., 2001) External EGR can

reduce in-cylinder oxygen concentration and cylinder

temper-ature due to its high heat capacity; as a result, NOx emission

can be reduced In addition, ignition timing is delayed so

that fuel-air mixing time is increased and local fuel-rich

zones are reduced; hence, soot emission can be reduced

However, external EGR also has some disadvantages For

example, EGR quantity is hard to control in transient mode;

it cannot promote in-cylinder mixing; finally, it takes part

of the intake air volume and affects air charge efficiency

In this paper, the authors used fuel injection during the

negative valve overlap period to realize diesel HCCI

com-bustion (Shi et al., 2005, 2007) In order to control the

combustion, CO2 was used to replace external EGR and

was directly injected into the engine cylinder by using an

air injection system Benefits of using CO2 in-cylinder

injec-tion are as follows: firstly, CO2 injection quantity can be

adjusted to satisfy the EGR requirement; secondly, by

ad-justing CO2 injection parameters (injection timing, pressure,

quantity), the CO2 jet adds extra momentum in the

cylinder; finally, the air charge efficiency is not affected if

the injection timing occurs after intake valve close timing

(IVC) In this paper, the effects of CO2 injection parameters

on HCCI combustion and emission characteristics were

investigated

2 EXPERIMENTAL APPARATUS AND

PROCEDURE

2.1 Test Engine

The test engine is a four-valve, single cylinder, naturally

aspirated DI diesel engine, and its specifications are shown

in Table 1 The fuel injection system is DENSO ECD-U2

common rail injection system The valve train is a Variable

Valve Timing (VVT) system In the experiments, the VVT

system was adjusted so that the engine had a valve overlap

of −30oCA, and a gas injection system was used to achieve

CO2 injection

In the experiments, AVL DiGAS 4000 light was used to

measure NOx by using an electrochemistry method and HC

and CO emissions by the nondispersive infrared (NDIR)

method Total HC emissions cannot be accurately

measur-ed by NDIR method; however, the HC emission trend

should be correct AVL 439 Opacimeter was used to

mea-sure smoke opacity

2.2 Gas Injection System

A gas injection system was used to inject CO2 into thecylinder It is composed of a high-resolution electronicscale (precision of 1 g, range of 0~100 kg), a gas bottle, apressure regulator, and a gas injector and its drive box The solenoid activated gas injector, manufactured byGuizhou Honglin Ltd, was installed in the cylinder head Itcomprises a solenoid valve and an injector body A checkvalve was installed in the injector body to prevent reverseflow of cylinder gas The nozzle tip of the gas injector has

a single hole, 4 mm in diameter, oriented 23o downwardinto the combustion chamber, as shown in Figure 1.The high-resolution electronic scale was used to mea-sure the mass reduction of the gas bottle in three minutesand then calculate cyclic gas injection quantity

Detailed information about the gas injection system andgas injection calibration can be obtained from a study by

Qu (Qu et al., 2008)

2.3 Experimental Procedure

In this paper, premixed diesel-air mixture was prepared byinjecting diesel early during the negative valve overlap period(see Figure 2) This early injected diesel can fully utilizethe heat of the internal EGR and piston to evaporate.Furthermore, the wall-wetting caused by over-penetration

of fuel can be avoided At the same time, to control HCCIcombustion, CO2 was injected directly into the cylinder.The fuel injection timing was studied first, followed by the

Table 1 Test engine specifications

Engine type 4-valve, DI, single cylinder

Displacement vol 2.15 LCompression ratio 14.8Combustion chamber ω typeFuel injection system Common Rail Injection SystemFuel injector nozzle 7×0.2 mm (Radial 6 holes)

Spray Angle 150o+Axial 1 hole)

EVC = 345oCA ATDCIVO =−345oCA ATDC

Figure 1 Cross section view of combustion chamber

EFFECT OF DIRECT IN-CYLINDER CO 2 INJECTION ON HCCI COMBUSTION 531

effects of CO2 injection parameters, including injection

timing, injection quantity, and injection pressure The engine

was run at 1300 r/min in all experiments

3 EXPERIMENTAL RESULTS AND ANALYSIS

3.1 Effect of Fuel Injection Timing

The effects of fuel injection timing on HCCI combustion

and emission characteristics for engine speed of 1300 r/min

and cyclic fuel injection quantity 0.057 g are shown in

Figure 3 and Figure 4, respectively Figure 3 indicates that

fuel injection timing (IT) has a large effect on cylinder

pressure and heat release rate (HRR) As the fuel injection

timing is delayed, maximum cylinder pressure and

maxi-mum heat release rate decrease Figure 4 reveals that, as the

injection timing is delayed, NOx decreases, HC and CO

increase, and smoke opacity remains unchanged In

addi-tion, HC and CO emissions do not significantly change

before injection timing of −365oCA but rapidly increase

afterwards This can be explained by the fact that late

injected fuel cannot sufficiently use the heat of internal

EGR, which decreases the vaporization ratio of fuel at

ignition timing and reduces combustion efficiency

The choice of fuel injection timing during negative valve

overlap must satisfy two criteria First, fuel injection timing

should advance to make use of heat of internal EGR

Second, injection timing should not be too early in order toavoid early ignition and excessively rapid combustion rate

In the following experiments, fuel injection timing wasfixed at −365oCA ATDC

Since the amount of NOx emission indicates the cylinder combustion condition, the authors propose to use

in-NOx emission as an indicator of upper load limit When theload increases to a certain level, auto-ignition occurs tooearly, such that the rapid heat release results in a highpressure rise rate, high peak cylinder pressure, and, hence,high mean gas temperature The resulted high temperatureleads to a high NOx formation rate and high NOx emission

In this study, the indicated mean effective pressure (IMEP)

at which NOx emission exceeded 100 ppm was taken as theupper load limit For engine speed of 1300 r/min and fuelinjection timing at −365oCA ATDC, the upper load limit isIMEP = 4.5 bar

3.2 Effect of CO2 Injection Timing on HCCI

To investigate CO2 injection timing on HCCI combustion,several cases in addition to the baseline case (without CO2

injection, n = 1300 r/min, cyclic fuel injection quantity 0.062

g, IMEP = 4.7 bar) were run At these cases, CO2 injectionpressure was 2 MPa, and gas injection pulse width was 4

ms Because the gas injection pressure was high whilecylinder pressure was low during the injection process, jetflow out of the gas injector remained choked throughoutthe injection process Therefore, the cyclic gas injectionquantity remained at a constant value, 0.092 g The gasinjection timing varied every 40°CA between −320oCAATDC and −80°CA ATDC (Figure 2)

CO2 injection timing determines its concentration bution at the end of a compression stroke An early inject-ing timing leads to a long mixing time with cylinder air and

distri-a reldistri-atively even CO2 concentration distribution; in trast, late injection timing leads to uneven distribution withlocally rich or poor CO2 concentration zones; hence,

con-Figure 2 Fuel injection and CO2 injection mode

Figure 3 Effects of fuel injection timing on HCCI

com-bustion

Figure 4 Effects of fuel injection timing on HCCIemission

heterogeneity is large Different CO2 distribution causes

different temperature distribution and affects local fuel

vaporization

Figure 5 shows the effects of CO2 injection timing on

cylinder pressure, mean gas temperature (T), pressure rise

rate (dp/dϕ), and heat release rate With an advance of CO2

injection timing, peak pressure decreases, and its

corre-sponding crank angle is delayed; mean gas temperature

decreases; both low and high temperature reactions are

delayed; pressure rise rate slightly decreases This can be

explained by the fact that as CO2 injection timing advances,

its concentration distribution becomes more even; hence, it

occupies a larger part of cylinder volume so that gas

temperature and pressure rise in these areas are reduced

due to the high heat capacity of CO2 Since HCCI ignition

is controlled by chemical kinetics, once the intake valve

closes and pressure-temperature-composition history

dur-ing the compression stroke determines the start of low

temperature heat release, advancing CO2 injection timing

can delay the start of low temperature reaction and, hence,

the start of high temperature reaction In addition, a

de-crease in local oxygen concentration reduces combustion

rate, resulting in low peak cylinder pressure and maximum

pressure rise rate A decrease of mean gas temperature is

not only related to lowered combustion rate and the highheat capacity of CO2, but also, to a certain extent, related tothe low temperature of injected CO2

Figure 6 illustrates the effects of CO2 injection timing onHCCI emission It can be seen that CO2 in-cylinder injec-tion can effectively reduce NOx emission With an advance

of CO2 injection timing, NOx rapidly decreases until −160

oCA; at this injection timing, NOx is below 100 ppm NOx

reduces with further advancing injection timing and reachesits lowest level at −240oCA ATDC, reduced by 91%relative to the baseline case Two reasons are considered to

be responsible for the reduction of NOx One is the tion of mean gas temperature The other is that, if CO2 isinjected before IVC, average in-cylinder oxygen concent-ration decreases Both reasons are beneficial for restraining

reduc-NOx formation

Smoke opacity slightly increases with CO2 injection due

to the reduction of cylinder temperature, which is mental to soot oxidation HC and CO also slightly increasedue to the decrease of temperature and oxygen concent-ration, which are detrimental to the oxidation reaction.3.3 Effect of CO2 Injection Quantity on HCCIFigure 7 shows the effects of CO2 cyclic injection quantity

detri-on cylinder pressure and heat release rate at CO2 injectiontiming of −240oCA ATDC With the increase of CO2 cyclicinjection quantity, peak cylinder pressure decreases, and itscorresponding crank angle is delayed; the starts of bothLTR and HTR are delayed This is because the high heatcapacity of CO2 is beneficial to lower temperature at theend of a compression stroke, and, hence, ignition delay isprolonged

Figure 8 illustrates the effects of CO2 injection quantity

on combustion characteristics With the increase of CO2

cyclic injection quantity, peak mean gas temperature (Tmax)and peak pressure rise rate ((dp/dϕ)max) decrease, and SOC

is delayed Compared to the baseline case, SOC is delayed

Figure 5 Effects of CO2 injection timing on HCCI

com-bustion

Figure 6 Effects of CO2 injection timing on HCCIemissions

EFFECT OF DIRECT IN-CYLINDER CO 2 INJECTION ON HCCI COMBUSTION 533

by 8oCA with CO2 cyclic injection quantity of 0.14 g Since

combustion phase delay effectively reduces compression

work, indicated thermal efficiency (η it) slightly increases

Figure 9 shows the effects of CO2 injection quantity on

emission levels It can be seen that increasing CO2 injection

quantity can dramatically reduce NOx emission With an

injection quantity of 0.14 g, NOx emission is reduced by

97%, from baseline’s 600 ppm to 19 ppm Lowered oxygen

concentration and combustion temperature are considered

to be responsible for this Soot emission depends on the

competition between its formation rate and oxidization

rate Lowered oxygen concentration tends to increase soot

formation rate; increased heat capacity and reduced

temper-ature tend to hinder soot oxidization rate As a result,

smoke opacity increases with the increase of CO2 injection

quantity HC and CO have a similar tendency

3.4 Effect of CO2 Injection Pressure on HCCI

To investigate the effect of CO2 injection pressure on HCCI

combustion, three injection pressures (1 MPa, 2 MPa, 3

MPa) were considered In all three cases, CO2 cyclic tion quantity was maintained at 0.092 g by adjusting injec-tion pulse width Figure 10 shows the effect of injectionpressure on emissions It can be seen that higher injectionpressure results in lower NOx and smoke opacity at mostinjection timings, while HC and CO rarely vary with pre-ssure change

injec-To explain these results, two assumptions are made First,

we assume that gas flow through the pressure regulator is

an adiabatic reversible process The pressure and ature before the regulator (gas bottle) are taken to be p 0 and

temper-T 0, respectively; the pressure and temperature after theregulator are p 1 and T 1, respectively Because the Joule-Thompson coefficient α h of CO2 is positive at room temper-ature as in Equation (1), larger injection pressure p 1 results

in larger temperature T 1

(1)Second, we assume that the gas injector can be simpli-fied as a convergent nozzle For ideal gas, the critical

α h = ∂T⎝⎛ -∂p⎠⎞ T 1 – T 0

p 1 – p 0

-> 0

≈

Figure 7 Effects of CO2 cyclic injection quantity on

cylinder pressure and heat release rate

Figure 8 Effects of CO2 cyclic injection quantity on HCCI

values of nozzle velocity Vcr, mass flow rate ,

momen-tum flux Mom flux, and total momentum Mom total are shown

in equations (2)~(5)

(2) (3)(4) (5)where k is the adiabatic coefficient, R is the gas constant of

ideal gas, A is the nozzle area, and ∆ m is the total injection

quantity Equations (2)~(5) indicate that critical velocity

and total momentum are only related to injected gas

temperature T 1, and momentum flux only relates to injected

gas pressure p 1

To simplify the analysis, we assume that equations (2)~

(5) are applicable to CO2 Under the second assumption,

the larger the injection pressure, the larger the momentum

flux Since larger injection pressure p 1 results in larger

temperature T 1, and larger T 1 results in larger total

momen-tum, then larger injection pressure also results in larger

total momentum

A jet with larger momentum flux and momentum

pro-vides stronger disturbance to the air-fuel mixture, resulting

in a more homogeneous mixture; therefore, larger injection

pressure leads to lower NOx emission and smoke opacity

Moreover, with the same CO2 injection quantity, the largest

injection pressure will accelerate the injection process and

shorten the injection duration, which has a similar effect as

advancing injection timing This also contributes to

emi-ssion reduction

Figure 11 shows the cylinder pressure and the heat

release rate at different injection pressures (1 MPa, 2 MPa,

3 MPa) All three cases have the same CO2 injection timing

(−200oCA ATDC) As shown in the figure, increasing

injection pressure delays SOC The trend is similar to that

of advancing injection timing

4 FUTURE WORK

In this paper, diesel HCCI combustion was realized byinjecting diesel during the negative valve overlap period.The injected fuel utilized the heat of internal EGR andpiston to vaporize, forming a homogeneous mixture beforeignition By analyzing the effect of fuel injection timing,the authors proposed to use −365oCA ATDC as the fuelinjection timing In the subsequent three sections, CO2

injection timing, quantity, and pressure were independentlystudied at a specific load condition (baseline case, IMEP =4.7 bar, NOx emission of 600 ppm) The results revealedthat either advancing injection timing or increasing injec-tion quantity can retard the start of combustion and reduce

NOx emission, while increasing injection pressure ced cylinder mixing and had an effect similar to advancinginjection timing

enhan-Since CO2 in-cylinder injection not only has the ages of external EGR but also its own merits, such as extradisturbance to the air fuel mixture and quick response, itseems that it is a promising method to control HCCIcombustion However, more work needs to be done toreach this conclusion Future work includes the followingaspects: (1) a study of the effects of injection direction; (2)

advant-a study of the interadvant-actions between CO2 jet flow and cylinder air flow by CFD tools, and (3) expanding HCCIupper load limit

in-5 CONCLUSION

(1) The choice of fuel injection timing during negativevalve overlap needs to satisfy two criteria First, fuelinjection timing should advance to maximize the utili-zation of heat of internal EGR Second, injectiontiming should not be premature to avoid early ignitionand rough combustion

(2) With the advance of CO2 injection timing, peak cylinderpressure, mean gas temperature, and pressure rise ratedecrease; starts of LTR and HTR are delayed With

CO2 injection pressure at 2 MPa and cyclic injectionquantity at 0.092 g, NOx reaches its lowest level at CO2

injection timing of −240oCA ATDC, reduced by 91%compared to the baseline case (IMEP = 4.7 bar, NOx

emission of 600 ppm)

(3) With the increase of CO2 injection quantity, peak meangas temperature and peak pressure decrease and SOC

is delayed With cyclic CO2 injection quantity of 0.14

g and injection pressure at 2 MPa, SOC is delayed

by 8oCA, indicated thermal efficiency is improved, and

NOx is reduced by 97% compared to the baselinecase

(4) CO2 injection pressure determines the extent of bance on the cylinder mixture High injection pressureresults in large momentum flux and total momentum,and hence, it is useful to increase mixture homogeneity,resulting in reduced NOx and smoke emission

p 1

∝ Mom total = ∆mV cr ∝ T 1

Figure 11 Effects of injection pressure on cylinder pressure

and heat release rate

EFFECT OF DIRECT IN-CYLINDER CO 2 INJECTION ON HCCI COMBUSTION 535

ACKNOWLEDGEMENTS− The authors wish to acknowledge

the financial support of National Key Fundamental R&D

Programs (973 projects, 2007CB210003) and National Nature

Science Foundation (Grant No 50406016).

REFERENCES

Choi, C Y and Foster, D E (1995) In cylinder augmented

mixing through controlled gaseous jet injection SAE

Paper No. 952358

Hardy, W L and Reitz, R D (2006) A study of the effects

of high EGR, high equivalence ration, and mixing time

on emissions levels in a heavy-duty diesel engine for

PCCI combustion SAE Paper No. 2006-01-0026

Kamimoto, T., Osako, S and Matsuoka, S (1983) An air

cell DI diesel engine and its soot emission characteristics

SAE Paper No. 831297

Kimura, S., Aoki, O., Kitahara, Y and Aiyoshizawa, E

(2001) Ultra-clean combustion technology combining a

low-temperature and premixed combustion concept for

meeting future emission standards SAE Paper No

2001-01-0200

Kitamura, T., Ito, T., Senda, J and Fujimoto, H (2002)

Mechanism of smokeless diesel combustion with

oxy-genated fuels based on the dependence of the equivalence

ration and temperature on soot particle formation Int J.

Engine Research 3, 4

Kurtz, E M and Foster, D E (1998) Exploring the limits

of improving di diesel emissions by increasing in-cylinder

mixing SAE Paper No. 982677

Kurtz, E M., Mather, D K and Foster, D E (2000) meters that affect the impact of auxiliary gas injection in

Para-a DI diesel engine SAE Paper No. 2000-01-0233.Nagano, S., Kawazoe, H and Ohsawa, K (1991) Reduc-tion of soot emission by air-jet turbulence in a DI dieselengine SAE Paper No. 912353

Okude, K., Mori, K., Shiino, S and Moriya, T (2004) mixed compression ignition (PCI) combustion for simu-ltaneous reduction on NOx and soot in diesel engine

Pre-SAE Paper No. 2004-01-1907

Onishi, S., Jo, S H., Shoda, K., Jo, P D and Kato, S.(1979) Active thermo-atmospheric combustion (ATAC)

- A new combustion process for internal combustionengines SAE Paper No. 790501

Peng, Z., Zhao, H and Ladommatos, N (2003) Effects ofair/fuel ratios and EGR rates on HCCI combustion of n-heptane, a diesel type fuel SAE Paper No. 2003-01-0747

Qu, S., Deng, K Y., Cui, Y and Shi, L (2008) Effects ofcarbon dioxide in-cylinder injection on premixed chargecompression ignition combustion J Automobile Engi- neering, Proc IMechE, Part D,222, 8, 1501−1511.Shi, L., Deng, K and Cui, Y (2005) Study of diesel-fueledHCCI combustion by in-cylinder early fuel injection andnegative valve overlap J Automobile Engineering, Proc IMechE, Part D, 219(D10), 1193−1201

Shi, L., Deng, K and Cui, Y (2007) Combustion stability

of diesel-fueled HCCI Int J Automotive Technology 8,

4, 395−402

EFFECTS OF GASOLINE, DIESEL, LPG, AND LOW-CARBON FUELS AND VARIOUS CERTIFICATION MODES ON NANOPARTICLE EMISSION CHARACTERISTICS IN LIGHT-DUTY VEHICLES

C L MYUNG 1) , H LEE 1) , K CHOI 1) , Y J LEE 2) and S PARK 1)*

1)School of Mechanical Engineering, Korea University, Seoul 136-701, Korea

2)Korea Institute of Energy Research, 102 Gageong-no, Yuseong-gu, Daejeon 305-343, Korea

(Received 4 November 2008; Revised 16 March 2009)

ABSTRACT− This study was focused on experimental comparisons of the effects of various vehicle certification modes on particle emission characteristics of light-duty vehicles with gasoline, diesel, LPG, and low-carbon fuels such as bio-diesel, bio- ethanol, and compressed natural gas, respectively The particulate matter from various fueled vehicles was analyzed with the golden particle measurement system recommended by the particle measurement programme, which consists of CVS, a particle number counter, and particle number diluters To verify particle number and size distribution characteristics, various vehicle emission certification modes such as NEDC, FTP-75, and HWFET were compared to evaluate particle formation with both CPC and DMS500 The formation of particles was highly dependent on vehicle speed and load conditions for each mode In particular, the particle numbers of conventional fuels and low-carbon fuels sharply increased during cold start, fast transient acceleration, and high-load operation phases of the vehicle emission tests A diesel vehicle fitted with a particulate filter showed substantial reduction of particulate matter with a number concentration equivalent to gasoline and LPG fuel Moreover, bio-fuels and natural gas have the potential to reduce the particulate emissions with the help of clean combustion and low-carbon fuel quality compared to non-DPF diesel-fueled vehicles.

KEY WORDS : Particulate matter, Nanoparticles, Diesel particulate filter, Differential mobility spectrometer, Condensation particle counter, Low-carbon fuels

1 INTRODUCTION

Diesel-powered engines have advantages of increased

engine power output, fuel economy, and higher durability

than spark ignition engines In addition, they can reduce

emissions such as hydrocarbons and carbon monoxide

Diesel engines are widely used in heavy-duty trucks, buses,

engine generators, etc., as they have fewer penalties in

performance and emissions In spite of the many

advant-ages, the emissions of smoke and particulate matter (PM)

from heavy-duty engines are a big drawback and are thus

the focus of many environmental researchers From the

viewpoint of health, PM emitted from diesel engines causes

adverse health effects, and recent studies have announced

that PM in the atmosphere is an important factor in

mortality and morbidity (Dockery et al., 1993; Giechaskiel

et al., 2007; Hagena et al., 2006; Ostro, 1984; Pope et al.,

1992; Takeda et al., 1995; Vaaraslahti et al., 2005)

In addition to diesel particles, conventional gasoline and

low-carbon fuels such as liquefied petroleum gas (LPG),

compressed natural gas (CNG), and various bio-fuels emit

a considerable amount of nanoparticles during the cold

start phase and at high-speed operating conditions; thus, theparticle formation mechanism of spark ignition engineshave been investigated (Choi et al., 2006; Kayes andHochgreb, 1999; Ristovski et al., 2000)

PM has been emphasized as a toxic air contaminant(TAC) by the California air resources board (CARB).Moreover, the developed countries have been focusing onthe effects of a variety of airborne particulates on healthrisks Current legislative exhaust emissions standards restrictparticle emission in terms of the total mass discharged perkilometer traveled Regulations based on total mass are aneffective way to control large particles; however, fineparticles contribute little to the total mass of particulatematter emissions (Andersson et al., 2001 and 2004)

In this context, the international particle measurementprogramme (PMP) has been developing a new particle sizemeasurement technique to complement or replace mass-based PM measurement procedures Final inter-laboratorycorrelation exercise (ILCE) results on particle number forLDV showed that particle number concentrations emittedfrom non-diesel particulate filter (DPF) diesel-fueled vehicles(E+13 particles/km) were much higher than from multi-point injection (MPI) gasoline engines (E+11 particles/km)and slightly higher than from gasoline direct injection engines

*Corresponding author. e-mail: spark@korea.ac.kr

538 C L MYUNG et al.

(E+12 particles/km), but particle numbers with DPF fitted

vehicles showed results equivalent to gasoline engines

(Andersson et al., 2007; Lee et al., 2008; Roberto et al.,

2007)

In diesel vehicle emissions, PM consists of tiny solid

particles and liquid droplets ranging from a few

nano-meters to around one micrometer in diameter (below 1,000

nm) PM size distributions are generally classified as

tri-modal The three modes are the nucleation mode,

accumu-lation mode, and coarse mode The nucleation mode is

typically composed of nanoparticles in the 5~50 nm

dia-meter range This mode consists of volatile organic and

sulfur compounds formed during the exhaust dilution and

cooling process The accumulation mode ranges from 50 to

1,000 nm and usually consists of particles that have been

deposited on cylinder walls and exhaust system surfaces

Finally, the coarse mode, rarely emitted in internal

com-bustion engines, is composed of particles with diameters

greater than 1,000 nm (Kittelson, 1998)

The goal of this research is to verify particle number and

size distribution characteristics under various vehicle

certi-fication modes such as the new European driving cycle

(NEDC), federal test procedure (FTP)-75, and the highway

fuel economy test (HWFET) modes using conventional

fuels, including low-carbon fuels that are widely used in

Korea automotive markets

2 EXPERIMENTAL APPARATUS AND

METHOD

2.1 Test Fuels and Vehicle Descriptions

Figure 1 shows the schematic diagram of the vehicle

experimental apparatus used to analyze the particle number

concentration and particle mass under the NEDC, FTP-75,

and HWFET modes To minimize fuel variation during test

periods, gasoline and diesel fuels were supplied from the

same filling station with one-batch preparation Summer

LPG and CNG for urban buses were used for gas-fueled

vehicles In the case of the 2.0 liter diesel engine, advanced

DPF meeting the EURO 4 emission regulation was

equipp-ed In addition, a 2.5 liter diesel engine run on 50% of

bio-diesel fuel was tested Additionally, a retrofitted 2.4 liter

bi-fueled CNG vehicle that can automatically switch between

gasoline and natural gas fuel was also tested The test

procedure for the bi-fueled CNG vehicle was as follows

To assess each fuel effect on particle formation in the CNG

vehicle, the gasoline fuel mode was tested first Then, the

natural gas mode was selected using the fuel selection

switch In this condition, natural gas was automatically

changed from gasoline during the NEDC mode when the

engine coolant temperature reached a target value In the

case of the ethanol flexible fuel vehicle (FFV), the ethanol

content was varied from gasoline to E85 (85% ethanol +

15% gasoline) To save time, only the NEDC test mode

was used for low-carbon fuels (FFV, bio-diesel) Test fuel

properties and vehicle specifications are summarized in

Table 1, Table 2, and Table 3

2.2 Particle Analyzer and Sampling SystemThe flow rate of the diluted exhaust gas through the CVStunnel was 20 m3/min at standard reference conditions (i.e.,20ºC and 1 bar) The primary dilution air was passedthrough a high-efficiency particulate air (HEPA) filter tominimize the particle effect of the background level of anemission facility

A sample probe for particles was fitted near the centerline in the dilution tunnel, and a cyclone was used as a pre-classifier to remove the particles with diameters greaterthan 2.5 µm in the CVS tunnel The number of particlesemitted from the test vehicle was counted using the goldenparticle measurement system (GPMS) which is recom-mended by PMP

Figure 2 represents the GPMS and particle mass system

Figure 1 Schematic diagram of vehicle experimentalsystem

Table 1 Properties of ethanol blended gasoline fuel

Table 2 Properties of bio-diesel fuel

components The volatile particle remover (VPR) comprises

the first particle number diluter (PND1), an evaporation

tube (ET), and a PND2 The PND1 is a rotating disc diluter

(MD19-2E) with the hot dilution set at 150°C and HEPA

filtered dilution air After the first diluter, the sample was

further divided into two flows The flow was conducted to

the ET held at a constant temperature of 300°C A 3010D

condensation particle counter (CPC) manufactured by TSI

was used to measure the time-resolved particle emission

number concentrations at NEDC, FTP-75, and HWFET

modes

As well as a CPC, a fast particulate spectrometer

(DMS500) was installed at the tailpipe location to analyzethe particle size distribution of gasoline, diesel with orwithout DPF, LPG, and bio-diesel-fueled vehicles at a high-speed driving condition (120 km/h)

Figure 3 shows the principle of the DMS500, whichprovides particle number and size distributions between 5

nm and 1,000 nm The particles were charged as theypassed through charger, and the charged particles landed in

a ring as function of charge and aerodynamic drag (FastParticulate Spectrometer user manual, 2008) The responsetime of the fast particle analyzer was on the order of 100ms

2.3 Calculation Procedure of Total Particle Numbers inVehicle Tests

Total particle number (N) emissions for vehicle drivingmode were calculated by means of the following equation

by particle number measurement procedure of regulation

No 83 (TRANS-WP29-GRPE-48, 2004)

(1)

In equation 1, N (particles/km) is particle numberemission expressed in particles per kilometer, V mix is thevolume of the diluted exhaust gas in liters per test, C avgisthe average concentration of particles in diluted exhaustgas in particles per cubic centimeter, DR tot is the totaldilution factor of the diluter in the VPR, and d is thedistance corresponding to test mode in kilometers

3 VEHICLE EXPERIMENTAL RESULTS

3.1 Particle Size Distribution during High Speed DrivingCondition

In order to understand particle formation characteristicsduring steady vehicle speed (120 km/h) of LPG, diesel,gasoline, and bio-diesel vehicles, each vehicle was driven

on a chassis dynamometer after reaching the fully

warmed-up condition The size distribution was measured with theDMS500 The sampling probe was positioned between the

N = Vmix × C avg × DR tot × 10 3

d -

Table 3 Specifications of test vehicles

Vehicle/fuel type LPG Gasoline Diesel Diesel (Bio-ethanol)FFV dieselBio- Bi-fuel (Retrofit) (Gasoline+CNG)

Engine displacement 1,998 cc 1,998 cc 1,991 cc 1,493 cc 1,596 cc 2,497 cc 2,359 cc

540 C L MYUNG et al.

tailpipe and the CVS inlet point

To reduce water condensation in the sampling probe and

delivery pipe, which can critically influence particle

count-ing, electrically controlled heating tape was applied to

maintain a constant temperature above 150°C For stable

particle analysis, the dilution ratio of the DMS500 was

experimentally selected and set to 4:1 (dilution air:exhaust

gas) in gasoline, LPG, and DPF diesel vehicles and 200 to

1,000:1 in non-DPF diesel vehicles to evaluate particle

formation on the optimization condition and to prevent

saturation

Figure 4 shows the particle size and number

distribu-tions during constant vehicle speeds for the LPG-, line-, diesel-, and bio-diesel-fueled vehicles PM size distri-butions in an internal combustion engine are generallyclassified into the nucleation and accumulation modeswhich are distinguished by the particle diameter Thenucleation and accumulation mode have particles of dia-meters less than about 50 nm and from 50 nm to 1,000 nm,respectively; however, the boundary between nucleationand accumulation mode is variable In the case of LPG,nucleation mode below d p< 10 nm was mostly observed,and the maximum particle number concentration was7.0E+4 particles/cm3

gaso-The particle sizes were mainly distributed in the 10 nm <

d p < 50 nm range in the case of gasoline, and the numberconcentration slightly increased compared to the LPGvehicle An especially high particle number was measured

in non-DPF diesel vehicles, with an order of magnitude ofabout E+8 particles/cm3, while the particle sizes weredistributed in accumulation mode around 100 nm in size.Comparing BD0 with BD50, as the bio-diesel contentsincreased, particle number concentration levels were re-duced Moreover, the particle size distribution of the DPF-equipped diesel vehicle showed a similar tendency to thenon-DPF diesel vehicle The particle emission level drasti-cally decreased to E+5 particles/cm3, an order of magnitudeequivalent to those of the advanced LPG and gasolinevehicles The large concentration of the diesel DPF vehicle

at 120 km/h could be explained by noting that the hightemperature inside the particulate trap caused the naturalregeneration of particles during the high-speed operatingcondition

Figure 4 Particle size distribution and number

concent-ration characteristics with different fuels at high speed

driving condition (120 km/h)

Figure 5 Time-resolved particle number concentrations of NEDC mode

3.2 Comparison of Time-Resolved Particle Number

Con-centrations in NEDC, FTP-75, and HWFET Modes

Figure 5 shows the time-resolved particle emission traces

of LPG, gasoline, and diesel vehicles measured by the CPC

for the NEDC mode The levels of PM number emissions

showed a close relationship to the driving condition of the

vehicle testing modes and the fuels used Particle number

concentrations of LPG, gasoline, and DPF diesel vehicles

were the highest during the cold start phase, on the order of

E+3 particles/cm3 Particle emissions gradually decreased

after the first transient and remained below 10 particles/cm3

except at the 120 km/h accelerating condition

The particle numbers and masses were, respectively,

9.95E+10 particles/km and 0.002 g/km for LPG, 1.44E+11

particles/km and 0.004 g/km for gasoline, and 1.09E+11

particles/km and 0.003 g/km for DPF-diesel However, the

non-DPF diesel vehicle showed higher particle mass and

number concentration, with values of 3.06E+13 particles/

km and 0.022 g/km, respectively These levels do not meet

the EURO 5 proposed number and mass regulation

stand-ards of 6.0E+11 particles/km and 0.0045 g/km,

respective-ly In the case of diesel fuel, the DPF after-treatment system

has the potential to greatly reduce particle numbers and

mass emissions

Figure 6 presents the time-resolved particle emission

behaviors of FTP-75 mode for LPG, gasoline, and diesel

vehicles Particle formation increased similarly to the NEDC

mode during vehicle speed-up; however, the concentrations

were primarily emitted during the cold start phase because

the vehicle speed gradient was steeper than in the NEDC

mode The total particle number and mass of the FTP-75

mode showed similar levels compared to NEDC To verifythe effect of vehicle speed on particle formation, HWFETmode was also tested

Figure 7 shows the time-resolved particle size tion spectra of LPG, gasoline, and diesel-fueled vehiclesduring HWFET mode using the DMS500 From theparticle spectra of HWFET mode, the nucleation mode wasmostly emitted by the LPG and gasoline-fueled vehicles,while the accumulation mode was observed in diesel-fueled vehicles The order of particle number emissionswas gasoline, LPG, DPF diesel, and non-DPF diesel.3.3 Evaluation of Particle Emissions for Low-Carbon FuelsLow-carbon fuels such as bio-ethanol, bio-diesel, and naturalgas have the potential to reduce regulated emissions andcarbon dioxide in automotive fuels

distribu-In addition, the oxygen component in bio-fuels

improv-ed combustion characteristics and enhancimprov-ed the rimprov-eduction

of harmful emissions In this section, NEDC mode wasselected to compare the particle emissions from these fuels.Figure 8 shows a comparison of particle number con-centrations of the bio-fueled vehicles From the figure, theparticle number concentration of E85 fuel was substantiallyreduced from 2.14E+11 particles/km for gasoline to1.35E+11 particles/km Moreover, particle emission duringthe last 400 seconds from the NEDC cycle was very low inE85 fuel, ascribed to the clean combustion characteristics

542 C L MYUNG et al.

from 9.40E+13 particles/km for BD0 to 7.29E+13 particles/

km for BD50 fuel Although a significant reduction was

observed in bio-diesel fuels, DPF was required to meet the

particle number regulation of 6.0E+11 particles/km enacted

to improve air quality in urban areas Natural gas fuel is

mainly composed of methane, which can enhance clean

combustion The bi-fueled CNG vehicle described in

Section 2.1 was tested

From the test result, CNG mode emitted very few

particles, about 3.21E+10 particles/km Moreover, the high

particle concentration in gasoline mode was not observed

during the last acceleration mode with natural gas

oper-ation Finally, considering gasoline operation during the

start transient phase, about 250 seconds, the particle

number concentration of the CNG vehicle can be reduced

significantly Total particle emissions were decreased by an

order of magnitude when bi-fuel was used

Table 4 summarizes the test results of particle numbers

and mass concentrations for various vehicle test modes

and fuel types used in this research From the table,

non-DPF diesel and bio-diesel-fueled vehicles have difficulty

in meeting the particle number and mass standards of

EURO 5 emission regulations with current after-treatment

systems

It should be noted that advanced LPG and MPI

gasoline-fueled vehicles showed PM emissions similar to those of

DPF diesel and had enough tolerance for future emission

regulations

Finally, CNG is the cleanest fuel for particle emissions

in terms of the fuel characteristics of gaseous natural gas

4 CONCLUSIONS

The objective of this research was to verify the particle

number and size distribution characteristics under various

vehicle certification modes such as NEDC, FTP-75, and

HWFET with conventional fuels and low-carbon fuels,

which are widely used in Korean automotive markets

Based on these analyses, the following major conclusions

can be drawn:

(1) Particle emission characteristics during constant vehiclespeed show that the particle number of non-DPF dieselreached the orders of E+7 particles/cm3 to E+8particles/cm3, while LPG, gasoline, and diesel vehicleswith DPF reached the orders of E+4 particles/cm3 toE+5 particles/cm3 In case of gasoline- and LPG-fueledvehicles, nucleation mode (dp< 50 nm) was the maincomponent of particulate matter However, accumulationmode (dp> 50 nm) was mostly emitted from diesel-fueled vehicles

(2) All the test fuels emit PM during transient vehicleoperation, including cold start, heavy acceleration phase,and high speed The diesel non-DPF vehicle showshigher particle mass and number concentration; how-ever, the diesel DPF vehicle shows a particle levelcomparable to gasoline- and LPG-fueled vehicles.(3) The orders of magnitude of the total particle numberconcentrations under various test modes for gasoline,LPG, and diesel (w/DPF) were similar However, theparticle concentration of non-DPF diesel is E+13 parti-cles/km, a level that has difficulty in meeting the particlenumber emission regulations enacted by EURO 5 and 6.(4) Alternative fuels such as bio-ethanol, bio-diesel, andnatural gas have the potential to reduce particulatenumber emissions due to their oxygen content and low-carbon fuel characteristics The particle number concent-ration of E85 fuel was reduced by 37% compared togasoline Moreover, as the bio-diesel content was varied

to 50%, the particle number level was reduced by 22%.Finally, the CNG-fueled vehicle emitted the lowestparticle number of 3.21E+10 particles/km among thevarious fuels tested

ACKNOWLEDGEMENTS− This study was supported by the Korea Petroleum Assciation and the ECO STAR Project of the Korea Ministry of Environment.

Table 4 Comparison of particle numbers and mass concentrations in vehicle test modes

Fuel type

Test modes and particle emissions

Number(particles/km) (g/km)Mass (particles/km)Number (g/km)Mass (particles/km)Number

Bio-ethanol

Bi-fuel

(CNG) GasolineBi-fuel 1.26E+113.21E+10 0.0050.005 8.99E+101.59E+11 0.0020.002 1.06E+105.41E+11

544 C L MYUNG et al.

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Roberto, C., Volker, S., Rainer, V and Thorsten, B (2007).Measurement of nucleation and soot mode particleemission from a diesel passenger car in real world andlaboratory in situ dilution Atmo Environ., 41, 2125−

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of heavy-duty diesel exhaust nanoparticles Environ Sci Technol., 39, 8497−8504

EFFECT OF FUEL INJECTION STRATEGIES ON THE COMBUSTION

PROCESS IN A PFI BOOSTED SI ENGINE

S S MEROLA, P SEMENTA, C TORNATORE * and B M VAGLIECO

Istituto Motori-CNR, Via Marconi 8-80125-Naples, Italy

(Received 13 November 2008; Revised 19 January 2009)

ABSTRACT− A low-cost solution based on fuel injection strategies was investigated to optimize the combustion process in

a boosted port fuel injection spark ignition (PFI SI) engine The goal was to reduce the fuel consumption and pollutant emissions while maintaining performance The effect of fuel injection was analyzed for the closed and open valve conditions, and the multiple injection strategies (MIS) based on double and triple fuel injection in the open-valve condition The tests were performed on an optical accessible single-cylinder PFI SI engine equipped with an external boost device The engine was operated at full load and with a stoichiometric ratio equivalent to that of commercial gasolines Optical techniques based on 2D-digital imaging were used to follow the flame propagation from the flame kernel to late combustion phase In particular, the diffusion-controlled flames near the valves and cylinder walls, due to fuel deposition, were studied In these conditions, the presence of soot was measured by two-color pyrometry, and correlated with engine parameters and exhaust emissions measured by conventional methods The open valve fuel injection strategies demonstrated better combustion process efficiency than the closed ones They provided very low soot levels in the combustion chamber and engine exhaust, and a reduction in specific fuel consumption The multiple injection strategies proved to be the best solution in terms of performance, soot concentration, and fuel consumption.

KEY WORDS : PFI SI Engine, Boosting, Multiple Fuel Injection Strategies, Fuel deposition, Optical diagnostics

1 INTRODUCTION

Currently, there are more than 260 million vehicles in-use

in Europe, and nearly all are powered by

reciprocating-piston internal-combustion engines burning

petroleum-derived hydrocarbon fuels Thus, the major challenge for

combustion scientists and engine-development engineers is

to optimize internal combustion engines to improve fuel

economy, reduce pollutant emissions, and provide

alter-native-fuel capabilities while maintaining outstanding

per-formance, durability, and reliability at an affordable price

With respect to gasoline-fuelled vehicles, several

techno-logical solutions have been explored; downsizing is

current-ly considered the best way to improve fuel economy, and

has a good cost to benefit ratio

Downsizing permits an increase in engine power and

torque without an increase in cylinder capacity This

pro-vides significant benefits in fuel consumption, as a result of

pumping losses reduction

Moreover, the lower engine capacity limits the

gases-to-wall heat transfer, because of a reduced internal surface

area, and has a shorter flame travel distance, due to a faster

combustion Moreover, it is possible to obtain lower

fric-tion losses due to smaller moving parts

Nevertheless, the car market requires invisible changes

from the driver’s point of view It is necessary to reduce theengine displacement volume while keeping torque andpower output constant The specific output performancemust then be increased by a ratio equal to the reduction inthe engine displacement Thus, an increase in the air andfuel content of the combustion chambers, from air boost-ing, is required Finally, the use of variable valve actuationallows for more enthalpy in the turbine, and thus a fasterturbocharged acceleration

The synergy between downsizing, turbo-charging ing), and valve actuation technologies provides opportuni-ties for gasoline engines to reduce the gap between fuelefficiency and CO2 emissions, compared to their dieselcounterparts (Alkidas, 2007; Heywood and Kasseris, 2007).All of these advanced technologies require the sub-stantial additional costs for engine hardware upgrades, andtarget performances seem hard to reach Thus, progress isstill needed in terms of in-cylinder combustion process andefficiency

(boost-In this work, a low-cost solution is proposed to optimizeboosted port fuel injection spark ignition (PFI SI) engines.Different fuel injection strategies are used to reduce thefuel consumption and pollutants emissions, while main-taining engine performance In particular, the effect ofinjection was investigated at closed and open valve condi-tions Moreover, multiple injection strategies (MIS) based

on double and triple fuel injections were tested in the

open-*Corresponding author. e-mail: c.tornatore@im.cnr.it

546 S S MEROLA, P SEMENTA, C TORNATORE and B M VAGLIECO

valve condition

These fuel injection strategies were considered in order

to improve combustion efficiency In fact, during the fuel

injection phase in PFI SI engines, thin films of liquid fuel

can form on the valves surface and cylinder walls (Costanzo

and Heywood, 2005; Gold et al., 2000; Shin et al., 1994;

Zhu et al., 2001) Successively, the fuel films interact with

the intake manifold and combustion chamber gas flow

(Meyer and Heywood, 1997) During the normal

combus-tion process, it is possible to achieve gas temperature and

mixture strength conditions that lead to fuel film ignition

This phenomenon can create diffusion-controlled flames

that can persist well after the normal combustion event

Just before the opening of the exhaust valves, these flames

produce soot that cannot be completely oxidized due to low

temperatures in the cylinder The chamber locations, where

the fuel-air mixture is too lean to burn, are a particular

concern for unburned hydrocarbon emissions Thus, fuel

film burning leads to an increase in smoke and

hydro-carbons emissions (Drake et al., 2003; Di Iorio et al., 2005;

Kim et al., 2005)

In this paper, the diffusion-controlled flame near the

valves and cylinder walls due to the burning of the fuel film

deposition were analyzed, and the soot distribution in the

combustion chamber was evaluated (Zhao and Ladommatos,

2001) by high resolution digital imaging and two-color

pyrometry The experiments were realized in a partially

transparent single-cylinder PFI SI engine with an external

boosting device The engine was equipped with a

four-valve real engine head, and fuelled with commercial

gaso-line In-cylinder optical investigations were correlated with

engine parameters and exhaust emissions, measured by

conventional methods

2 EXPERIMENTAL APPARATUS

2.1 Transparent Engine

An optically accessible single cylinder PFI SI engine was

used for the experiments It was equipped with a new

generation SI turbocharged engine-cylinder head with four

valves A ten-hole injector with 25o spray angle and 70 mm

SMD was employed Details about the engine are reported

in Table 1 Figure 1 shows a sketch of the bottom field of

view of the combustion chamber

The head has four valves and a centrally located spark

plug A quartz pressure transducer was flush-installed in

the region between intake-exhaust valves Combustion ssure measurements were performed for all the selectedoperating conditions The in-cylinder pressure, rate ofchemical energy release, and all related parameters wereevaluated, on an individual cycle basis and/or averagedacross 400 cycles, from the cylinder pressure data usingconventional interpretation models (Heywood, 1998)

pre-An external device controlled the intake air pressureover a range of 1000~2000 mbar, and the temperature over

a range of 290~340 K The engine piston was flat andtransparent through a quartz window (φ= 57 mm) Toreduce window contamination by lubricating oil, an elon-gated piston arrangement was used with unlubricatedTeflon-bronze composite piston rings in the optical section.2.2 Optical Setup for Optical Measurements

Figure 2 shows the experimental apparatus for opticalinvestigations During the combustion process, the lightpasses through a quartz window, located in the piston, and

is reflected toward the optical detection assembly by a 45o

inclined UV-visible mirror, located in bottom of the engine.2D soot flame visualization was obtained by a 12-bitdigital CCD color camera coupled with a 50 mm focal

Table 1 Specifications of the single cylinder boosted port

fuel injection (PFI) engine

length, f/3.8 Nikon lens The CCD had a 640×480 pixel

matrix with a pixel size of 9.9×9.9µm2 This optical

assessment allowed for a spatial resolution of around 100

µm/pixel The spectral range of the camera was 290~800

nm Spatial distribution of soot temperature and

concent-ration was obtained by the two color method The

soot-emission wavelengths were selected by edge filters More

details about this methodology are reported in (Zhao and

Ladommatos, 2001; User manual of AVL, 2003)

The camera and engine were synchronized by sending

the Crank Angle Encoder signal through a unit delay The

exposure time was fixed at 41.6 ms, which corresponds to a

0.5o crank angle (CA) at an engine speed of 2000 rpm The

camera was not a cycle resolved detector In this work,

each image was detected at a fixed crank angle for different

engine cycles The dwell time between two consecutive

images was 41.6 µs

2.3 Exhaust Measurements

Steady-state measurements of CO, CO2, O2, HC, and NOx

were acquired from the raw exhaust by AVL analyzers

CO, CO2, and HC were measured by Non-Dispersive

Infrared Detectors (NDIR); NOx and O2 were detected by

means of an electrochemical sensor An Opacimeter was

used to measure particulate mass concentration The

opaci-meter is a partial-flow system that measures visible light

attenuation (550 nm) from the exhaust gases By empirical

relations, it is possible to convert the opacity percentage to

particulate mass concentration (Mörsch and Sorsche, 2001)

3 RESULTS AND DISCUSSION

3.1 Engine Operating Conditions

All of the tests presented in this paper were carried out at

an engine speed of 2000 rpm at full load Absolute intake

air pressure and temperature were fixed at 1400 mbar and

323 K, respectively Commercial 95 octane gasoline was

used

Different fuel injection strategies were tested (Table 2).Initially, two single injections in closed-valve (CV1) andopen-valve (OV1) conditions were considered Then, theeffect of multiple injection strategies (MIS) was investi-gated; in particular, two double fuel injection strategieswere analyzed The OV2-360 had the highest dwell timebetween the two injections, and operated in an open-valvecondition The OV2-320 condition was characterized bythe minimum dwell time allowed by the mechanical inertia

of the injector Moreover, a triple injection strategy

OV3-365 was investigated Three injections were the highestnumber allowed by the mechanical inertia of the injector.For all test cases, the injection-duration was chosen tomaintain a stoichiometric equivalent ratio, as measured by

a lambda sensor at the engine exhaust The spark timingwas fixed to always operate in the Maximum Brake Torque(MBT) condition (Zhu et al., 2003) More details about theengine operating conditions are reported in Table 2.3.2 Single Injection Strategy

Figure 3 reports the pressure averaged curves for the threesingle injection strategies The related Indicated Mean Effec-tive Pressure (IMEP) and Coefficient of Variation (COVIMEP) are listed in Table 3 Even if the cycle-to-cycle vari-ation and heat transfer between the different components ofthe optical engine induced a thermal evolution and fluctu-ation of the maximum pressure signal, the IMEP and COVdemonstrated results that were comparable to those mea-sured for real multi-cylinder engines (Alkidas, 2007) More-over, these data show an improvement, in terms of stability,between the closed and open valve injection phases.This result matches with the decrease in CO emissionreported in Table 3 In fact, the level of CO emission in theexhaust of an IC engine varies with the fuel-air ratio Forfuel-rich mixtures, high CO concentrations in the exhaustemissions are generally observed (Sayi et al., 2007) Since,

in these experiments, the engine operated at stoichiometricfuel-air ratios, the CO emission in the exhaust was a result

Figure 2 Optical setup for digital imaging

Table 2 Engine operating conditions

Test label Number of injection Duration of injection

[CAD]

Start of injection ATDC[CAD]

−365

−275

−185

548 S S MEROLA, P SEMENTA, C TORNATORE and B M VAGLIECO

of the presence of locally rich zones in the chamber Thus,

it is possible to state that in the OV1-360 condition, a more

complete combustion took place, compared to the other

single injection strategies

The pressure measurements and exhaust characterization

provided only overall information, and did not allow forlocally resolved combustion process information (Zhao et

al., 2001) For local measurements, optical techniques aremore suited to give details about the thermo and fluiddynamic phenomena that occur in the combustion chamber.Previous optical investigations in comparable operatingconditions (Merola et al., 2008) showed that the firstevidence of the flame was detectable at 2 CAD after thespark timing (CAD ASOS) Then the flame front spreadwith a radial-like behavior for about 16 CAD

This result agrees with the optical data presented inFigures 4 and 5, which show digital images selected torepresent the flame propagation for single fuel injectionstrategies From 16 CAD ASOS, an asymmetry in the

Figure 3 Injection signal and pressure averaged over 400

consecutive cycles for the CV1, OV1-300, and OV1-360

conditions, respectively

Table 3 Indicated mean effective pressure, coefficient ofvariation, and CO and HC exhaust emissions for closed-valve and open-valve single injection strategies

test IMEP[bar] COV IMEP [%] [g/kWh]CO [g/kWh]HC

flame front shape was observed This was caused by the

fuel film deposited in the intake manifold In the closed

valve conditions, this film dripped on the intake valve

steam Then, it accumulated on the valve seat and was

drawn by gravity onto the valve head, where it remained as

a film, due to the surface tension (Costanzo and Heywood,2005) Once formed in the combustion chamber, the fuelfilm developed dynamically under the influence of gas flow(Heywood and Kasseris, 2007; Costanzo and Heywood,2005; Gold et al., 2000; Shin at al., 1994; Zhu et al., 2001).The heat exchange between the intake ports and surround-ing gas led to the fuel evaporation This effect influencedthe composition of the mixture and hence, the combustionprocess In particular, the liquid fuel vaporization canreduce the flame speed and complete flame propagation,creating locally rich-zones

In the open-valve condition, the injected fuel dropletswere carried by gas flow into the combustion chamber.Some stuck to the cylinder walls while some was deposited

on the piston surface These fuel depositions created locallyfuel-rich zones, millimeters size When the flame appro-ached these zones, several small ignition surfaces appeared

as bright spots, shown in Figure 5 (Merola and Vaglieco,2007; Merola et al., 2008)

The aforementioned processes induce a different air tofuel ratio distribution in the combustion chamber for theopen and closed valve injection conditions This can be thereason of the faster flame front spread in the open valveconditions

Figures 6 and 7 show the flame emissions for singleinjection strategies from 25 CAD ASOS until the opening

of the exhaust valves (155 CAD ATDC) When the flamefront interacted with the fuel deposits on the intake valvesand cylinder walls, diffusion-controlled flames were obser-ved (Merola and Vaglieco, 2007; Merola et al., 2007) Inthe closed valve condition (CV1-130), the highly intensediffusion-controlled flame was observed near the intakevalves, as expected (Figure 6) In the open-valve conditions(OV1-300 and OV1-360), the flame near the intake valveswas less intense, and a strong flame emission was detected

Figure 5 Flame emission detected in the chamber for the

open valve condition (OV1)

Figure 6 Flame emission detected in the late combustionphase for the closed valve condition (CV1-130)

550 S S MEROLA, P SEMENTA, C TORNATORE and B M VAGLIECO

near the cylinder walls

As reported in previous works, the spectra detected in the

late combustion phase for the closed-valve condition were

characterized by a strong continuous contribution that

increased with wavelength in the visible range This was

the typical soot emission (Zhao et al., 2001) Similar

spec-tral features were observed for the open-valve condition

near the chamber walls (Merola and Vaglieco, 2007)

In order to study the effect of the diffusion controlled

flame on soot concentration, the two-color pyrometry

technique was applied (Zhao et al., 2001; User manual of

AVL, 2003) Figure 8 shows the time evolution of soot

concentration, averaged over the combustion chamber

volume, for the single injection conditions investigated

The highest soot signal was detected for CV1 In this

condition, a more heterogeneous distribution of fuel

de-position in the combustion chamber is shown, compared to

OV1 conditions Due to the volatile species vaporization,

fuel-rich regions, with a high density of low-volatile

com-pounds, were created near the valves This confirms the

higher HC level measured at the exhaust in CV1 The high

HC concentration also promoted soot precursor formationand thus, the CV1 condition showed higher soot levelsduring the entire cycle Moreover, it is possible to observethat the soot reduction rate at the exhaust valve openingwas not sufficient to oxidize all of the particulates for CV1

To resume, the open-valve conditions showed not only adifferent spatial distribution of diffusion-controlled flamethan the closed-valve condition, but also a different timeevolution of soot and particulate matter in the combustionchamber

3.3 Multiple Injection Strategy

In order to investigate low-cost solutions for combustionprocess optimization in terms of fuel consumption and PMemission, some multiple injection strategies (MIS) wereconsidered in the open-valve condition Figure 9 reportsthe MIS pressure curves averaged over 400 cycles Therelated IMEP and COV are included in Table 4 MIS strate-gies yield the same performance as single injection strate-gies in terms of IMEP, and better stability, as confirmed byCOV reduction

To better understand the effect of MIS on the tion process, 2D digital imaging of the flame was perform-

combus-ed A selection of images detected in the late combustionphase is reported in Figure 10 For all multiple injectionstrategies, the luminosity of the flame in the combustionchamber was lower than the single injection ones More-over, the residual luminosity at the exhaust valve openingwas strongly reduced

Two-color pyrometry was applied to evaluate the MISsoot concentration, averaged over the combustion chambervolume The results are reported in Figure 11 It can benoted that the total soot formed in the combustion chamberdecreased when the number of injections increased, due tobetter fuel-air interaction in the intake manifold This effect

is lower in the OV2-320 than in the OV2-360, due to the

Figure 7 Flame emission detected in the late combustion

phase for the open valve condition (OV1)

Figure 8 Average soot concentration measured in the CV1and OV1 conditions

lower dwell time between the two injections Therefore, a

more intense flame with a greater area was detected The

balance between the fuel injection duration and dwell time

allowed for the best vaporization in the triple injection

strategy Thus, the lowest level of soot was formed

It should be noted that the optical engine lacks an

after-treatment device On the other hand, the three way catalystswere not able to remove soot particles, as reported in recentexperimental studies (Merola and Vaglieco, 2007) In fact,

a high number of nanometric carbonaceous particles can bedetected at the exhaust of PFI SI engines, not only duringcold start-up, but also at low speeds (Di Iorio et al., 2005;

Figure 9 Injection signal and pressure averaged over 400

consecutive cycles for multiple injection strategies (MIS)

Table 4 Indicated mean effective pressure, coefficient ofvariation, and gases exhaust emissions for multiple injec-tion strategies (MIS)

test IMEP[bar] COV IMEP[%] [g/kWh]CO [g/kWh]HC

552 S S MEROLA, P SEMENTA, C TORNATORE and B M VAGLIECO

Maricq et al., 2006) Thus, the reduction in fuel film

de-position burning becomes fundamental for the optimization

of new generation SI engines

The higher combustion process efficiency of the open

valve fuel injection strategies provided a strong reduction

in exhaust particulate matter emission, as reported Figure

12 The effect is particularly positive for MIS This result isinteresting, because it is coupled with the reduction inspecific fuel consumption (Figure 13) In fact, in the openvalve fuel injection strategies, an improvement of theBSFC in the range 2.0~3.5% was measured Moreover, themultiple injection strategies were characterized by thelowest soot levels in the combustion chamber and engineexhaust, and yielded the best blend of performance and fuelconsumption

ed with a four-valve head and external boost device Theengine worked at full load and stoichiometric equivalentratio

Optical techniques, based on 2D-digital imaging, wereused to follow the flame propagation in the combustionchamber In particular, the diffusion-controlled flame nearthe valves and cylinder walls, due to the different fuelinjection strategies, were investigated Two-color pyrometrywas employed to measure the soot temperature and con-centration In-cylinder optical investigations were correlat-

ed with the engine parameters and exhaust emissions,measured by conventional methods

The effect of the injection at closed and open valveconditions was studied Moreover, multiple injection strate-gies (MIS), based on double and triple fuel injection, weretested in the open-valve condition

The experiments demonstrated that the open valve fuelinjection strategies were characterized by higher com-bustion process efficiency than closed valve ones Theyprovided a strong reduction in particulate matter in thecombustion chamber and engine exhaust, and a substantialdecrease in specific fuel consumption The multiple injec-tion strategies confirmed their potential for optimizing theboosted PFI SI engines, showing the lowest soot level andfuel consumption

ACKNOWLEDGEMENT− The authors are grateful to Mr Carlo Rossi and Mr Bruno Sgammato for the assessment of the optical engine and for the support in the experimental activities REFERENCES

Alkidas, A C (2007) Combustion advancements in line engines Energy Conversion and Management,48,2751–2761

gaso-Costanzo, V S and Heywood, J B (2005) Mixture ration mechanisms in a port fuel injected engine SAE Paper No 2005-01-2080

prepa-Di Iorio, S., Merola, S S., Vaglieco, B M and Tornatore,

Figure 11 Average soot concentration measured in

multi-ple injection strategies (MIS)

Figure 12 Exhaust soot reduction for the open valve

conditions with respect to the closed valve condition,

evaluated as a percentage

Figure 13 Specific Fuel Consumption for the open valve

conditions with respect to the closed valve condition,

evaluated as a percentage

C (2005) Nanoparticles characterization at spark

igni-tion engine exhaust SAE Paper No. 2005-24-010

Drake, M C., Fansler, T D., Solomon, A S and Szekely,

G A Jr (2003) Piston fuel films as a source of smoke

and hydrocarbon emissions from a wall-controlled

spark-ignited direct-injection engine SAE Paper No.

2003-01-0547

Gold, M R., Arcoumanis, C., Whitelaw, J H., Gaade, J

and Wallace, S (2000) Mixture preparation strategies in

an optical four-valve port-injected gasoline engine Int.

J Engine Research 1, 1, 41−56

Heywood, J B and Kasseris, E (2007) Comparative

analysis of automotive powertrain choices for the next

25 years SAE Paper No 2007-01-1605

Heywood, J B (1998) Internal Combustion Engine

Fund-amentals. McGraw-Hill New York

Kim, H., Yoon, S and Lai, M.-C (2005) Study of

corre-lation between wetted fuel footprints on combustion

chamber walls and UBHC in engine start processes Int.

J Automotive Technology 6, 5, 437−444

Maricq, M., Xu, N and Chase, E (2006) Measuring

parti-culate mass emissions with the electrical low pressure

impactor J Aerosol Science and Technology, 40, 68−79

Merola, S S and Vaglieco, B M (2007) Optical

investi-gations of valve firing in PFI spark-ignition engine

Proc ECOS 2007 Conf., Paper No. ECOS07-M01–I,

105

Merola, S S., Sementa, P., Tornatore, C and Vaglieco, B

M (2008) Effect of injection phasing on valves and

chamber fuel deposition burning in a PFI boosted ignition engine SAE Paper No. 2008-01-0428

spark-Merola, S S., Sementa, P., Tornatore, C and Vaglieco, B

M (2007) Effect of fuel film deposition on combustionprocess in PFI SI engine J KONES Powertrain and Transport 14, 3, 395−402

Meyer, R and Heywood, J B (1997) Liquid fuel transportmechanisms into the cylinder of a firing port-injected SIengine during start up SAE Paper No. 970865 Mörsch, O and Sorsche, P (2001) Investigation of Alter- native Methods to Determine Particulate Mass Emissions.UNECE/WP.29/GRPE Report 2001

Sayi, C., Ertunc, H M., Hosoz, M., Kilicaslan, I andCanakci, M (2007) Performance and exhaust emissions

of a gasoline engine using artificial neural network

Applied Thermal Engineering, 27, 46–54

Shin, Y., Cheng, W K and Heywood, J B (1994) Liquidgasoline behaviour in the engine cylinder of a SI engine

SAE Paper No. 941872

User Manual of AVL (2003) AVL Graz Austria 2003.Zhao, H and Ladommatos, N (2001) Engine Combustion Instrumentation and Diagnostics SAE Int., Inc.Zhu, G G., Daniels, C H and Winkelman, J (2003) MBTtiming detection and its closed-loop control using in-cylinder pressure signal SAE Paper No 2003-01-3266.Zhu, G S., Reitz, R D., Xin, J and Takabayashi, T (2001).Modelling characteristics of gasoline wall films in theintake port of port fuel injection engines Int J Engine Research 2, 4, 231−248

International Journal of Automotive Technology , Vol 10, No 5, pp 555 − 560 (2009)

555

PERFORMANCE OF AIR CONDITIONERS WITH A GAS-LIQUID SEPARATION CONDENSER AND ONE-TANK LAMINATED

EVAPORATOR

S.-Y YOO 1) , D.-W LEE 2)* and M.-S JIE 3)

1)Department of Mechanical Design Engineering, Chungnam National University, Daejeon 305-764, Korea

2)Advanced Engineering Team, R & D Center, Halla Climate Control Corporation, 1689-1 Sinil-dong,

Daedeok-gu, Daejeon 306-230, Korea

3)Department of Automobile Engineering, Ajou Motor College, Chungnam 355-769, Korea

(Received 14 November 2008; Revised 10 March 2009)

ABSTRACT− In this study, a parallel flow condenser and laminated evaporator for an automotive air-conditioning system were modified to improve performance Gas-liquid separation type condensers, in which the condenser and receiver drier are integrated, and one-tank laminated type evaporators were developed, and their performances were investigated experimentally using HFC-134a Heat transfer characteristics in the condenser are examined by means of air temperature, air velocity entering the condenser and inlet pressure of the refrigerant; heat transfer characteristics in the evaporator are examined by means of air temperature, relative humidity, flow rate of air, outlet pressure of refrigerant and superheat Pressure drops for both evaporator and condenser are also measured, and correlations for pressure drop are derived for the condenser and evaporator, respectively Air velocity and mass flow rate of the refrigerant have a significant effect on the overall heat transfer coefficient, and flow pass

is not significantly influenced by the cooling capacity of the condenser The overall heat transfer coefficient of the evaporator increases as air flow rate, air temperature and relative humidity increases.

KEY WORDS : Gas-liquid separation condenser, One-tank laminated evaporator, Heat transfer coefficient, Pressure drop, Automotive air-conditioning system

1 INTRODUCTION

Automotive air conditioning systems are usually installed

in limited spaces; thus, compactness and improved

perfor-mance are required for successful designs of automotive

heat exchangers Early evaporators and condensers in

auto-motive air conditioning systems were fin-tube typed designs

Recently, evaporators have transformed from serpentine to

laminated types, and condensers have changed from

ser-pentine to parallel flow types

Parallel flow type condensers with louvered fins have

been developed to achieve higher performance, compared

to the fin-tube or serpentine type condensers (Sugihara and

Lukas, 1990) An experimental study on the heat transfer

and friction characteristics of air has been performed by

Sahnoun and Webb (1992) Webb et al (1995) conducted

different experiments on 32 copper-brass and 27 aluminum

heat exchangers, and established equations for heat transfer

and frictional coefficients Ali et al. (1995) suggested

improvements through pressure drop and heat transfer

modeling for parallel flow type condensers (i.e., separated

inlet, middle and outlet areas) It was reported that the

outlet area should be 30 to 60% of the inlet area to achievethe lowest pressure drop and higher heat transfer Masafumi(1994) established heat transfer characteristics for multichannel tube (4 to 30) condensers with different channelheights and widths Multi-channel tube condensers havesmaller hydraulic diameter tubes, with a reduced chargequantity of refrigerant, and can be produced through extru-sion (Lee et al., 1996a) These kinds of condensers arewidely used in automotive air conditioning systems Inaddition to this, Yamanaka et al (1997) introduced sub-cooled condensers, which increases condensing efficiency

by 5% and reduced power consumption by 10%, as a result

of an auxiliary cooling mechanism at the condenser outlet.Laminated type evaporators were introduced by Kurosawaand Noguchi (1987) Ohara and Takahashi (1988) reportedforced convection heat transfer and two phase flow charac-teristics using CFC-12 for laminated evaporators with atransverse slit shape Ohara et al (1990) also conducted anexperimental test to establish the relationship between pre-ssure drop and heat transfer characteristics at the smoothwall and transverse rib shape plate He found that heattransfer at the transverse rib is about three times higherthan at the smooth wall Kandlikar (1990) and Kaviany(1993) studied pressure drop and heat transfer characteri-

*Corresponding author. e-mail: dwlee@mail.hcc.co.kr

stics for slit shaped rib laminated plates with U-turn flow

and HFC-134a Kang (1995) carried out experiments on

elliptical dimpled laminated plates at various test

condi-tions, vapor quality, mass velocity of refrigerant and amount

of refrigerant oil

In this study, parallel flow condensers and laminated

evaporators are modified to improve performance

Gas-liquid separation type condensers, in which condenser and

receiver drier are integrated, and one-tank laminated type

evaporators have been developed Performances of the

gas-liquid separation type condenser and one-tank laminated

type evaporator for automotive air conditioning systems

are investigated experimentally using HFC-134a From

experimental data, such as air temperature, flow rate of air,

relative humidity, refrigerant pressure and temperature,

overall heat transfer coefficients are calculated for both

condenser and evaporator Pressure drop is then correlated

as a function of Reynolds number, based on refrigerant

mass flow rate and hydraulic diameter

2 EXPERIMENTAL APPARATUS AND

CONDITIONS

2.1 Condenser and Evaporator Calorimeter

The experimental apparatus (calorimeter), shown in Figure

1, consists of a test section, condenser control section,

evaporator control section, air heating system, air cooling

system and control panel The calorimeter has guide blades,

which allow for uniform air flow, and is insulated from the

environment Heat gain and loss by radiation are

com-pensated for by accurate measurements

The air heating system has a closed electric heater which

can heat the circulated air more regularly than an open type

electric heater

The air cooling system is controlled by a 30-HP

com-pressor, expansion valve and evaporator pressure regulator

Two oil separators are installed in series to minimize

per-formance error due to compressor oil Oil is circulated less

than 3% of refrigerant mass flow Temperature is measured

at the inlet and outlet of the experimental apparatus by

using a RTD sensor Brandt B-NZP1000 nozzles with 10

cm and 30 cm diameters are used to measure low and high

flow rate of air, respectively Next, 180oC steam is supplied

to the chamber to control humidity, and chilled mirror typehumidity sensors, which can calculate air enthalpy differ-ences, were used Flow rate of the refrigerant is measuredusing a mass flow meter

2.2 Experimental ConditionsFor gas-liquid separation type condensers, the heat transferand pressure drop phenomenon is somewhat different fromthose of conventional multi channel type condensers be-cause refrigerant flow through the condenser is divided intoupper and lower parts at the same time As shown in Table

1, experimental conditions, such as air velocity, air ature, and refrigerant pressure of the condenser inlet, areconsidered for air and refrigerant sides

temper-For laminated type evaporators, performance depends onthe number, length and angle of the dimple or rib Heattransfer and pressure drop characteristics are very sensitive

to these conditions In this study, to successfully design aslim laminated type evaporator with louver fin, experimentsare performed by changing evaporator inlet air temper-ature, relative humidity, air flow volume, evaporator outletrefrigerant pressure and superheat, as shown in Table 2

3 RESULTS AND DISCUSSION

3.1 CondenserThe measured pressure drop ∆P for the condenser isillustrated in Figure 2 Pressure drop increases as theReynolds number increases, in which Reynolds number isdefined by gas viscosity of refrigerant as follows,

(1)where G r is the mass flow rate, D h is the hydraulic diameter

of tubes, A is the refrigerant flow area, and µ g is theviscosity of the refrigerant gas Experimental results are

Re g = G r ⋅ D h

µ g ⋅ A -

Figure 1 Schematic diagram of evaporator and condenser

calorimeter

Table 1 Experimental conditions for the condenser.Air inlet temperature (oC) 30, 37Air inlet velocity (m/s) 2, 3, 4, 5Refrigerant inlet pressure (kPa) 1474, 1739, 1867Super-heated temperature (oC) 25

Table 2 Experimental conditions for the evaporator.Air inlet temperature (oC) 25, 27, 30Air inlet relative humidity (%) 50, 60, 70Air volume flow rate (m3/h) 300, 420, 500Refrigerant inlet pressure (kPa) 1621Refrigerant outlet pressure (kPa) 278, 297, 317, 356Super-heated temperature (oC) 5, 10

PERFORMANCE OF AIR CONDITIONERS WITH A GAS-LIQUID SEPARATION CONDENSER 557

correlated using the least-squares method, as shown in

Equation (2)

(2)The amount of heat transfer at the condenser can be

expressed in terms of LMTD (log mean temperature

difference), and overall heat transfer coefficient and area,

as in Equations (3) and (4),

(3) (4)

where U is the overall heat transfer coefficient, A is the heat

transfer area, and subscript ri, ro, ai and ao indicate

refrigerant inlet, refrigerant outlet, air inlet and air outlet

temperature, respectively The overall heat transfer

coeffi-cient of the condenser depends on geometric

characteri-stics, Reynolds number inside the tube and air velocity

Figure 3 shows the variation of the overall heat transfer

coefficient with the flow rate of refrigerant, for various air

velocities at superheat of 25oC and sub-cooling of 5oC The

overall heat transfer coefficient increases as refrigerantmass flow rate and air velocity increases Figure 4 showsthe variation of the overall heat transfer coefficient with airtemperature entering the condenser and refrigerant pressure

at condenser inlet The overall heat transfer coefficient has

a tendency to increase at low condenser inlet pressures forthe same refrigerant flow rate, because refrigerant vaporvelocity at condenser decreases as pressure, temperatureand density increases

Variation of condensing capacity with air velocity at thecondenser front area is shown in Figure 5 for various airtemperatures and refrigerant pressures at the condenserinlet Condensing capacity is high at lower air temper-atures, and inlet pressure and condensing capacity at 5 m/sair velocity is up to 40% higher than at 2 m/s

Figure 6 presents the condensing capacity and pressuredrop for various refrigerant flow passes inside the conden-ser at an air velocity of 5 m/s, superheat 25oC and sub-cooling 5oC The sub-cooling region at the condenser islimited to within 14 to 17% of the total area, superheatingregion by 36 to 42%, two-phase region of the upper con-denser by 30 to 33% and lower condenser by 11 to 17%.Cooling capacity is not significantly influenced by the

-Figure 2 Variation of pressure drop with Reynolds number

for the condenser

Figure 3 Variation of heat transfer coefficient with flow

rate of the refrigerant

Figure 4 Variation of heat transfer coefficient with inletpressure and air temperature

Figure 5 Variation of condensing capacity with air velocity

refrigerant flow pass, but pressure drop has a tendency to

increase as the superheat region becomes larger and

sub-cooling region becomes smaller The maximum difference

in pressure drop is about 12%

3.2 Evaporator

Variation of pressure drop with Reynolds number is

pre-sented in Figure 7, in which Reynolds number is defined by

(5)where A is the refrigerant flow area; the inlet refrigerant

viscosity is calculated from the refrigerant saturation vapor

viscosity coefficient µ g and vapor quality (x) at the

evapo-rator inlet

Pressure drop increases as Reynolds number increases,

and the slope is clearly different for superheat 5oC and

10oC At the same Reynolds number, as superheat (Tsh)

increases, pressure drop also increases Pressure drop ∆P is

correlated in terms of Reynolds number for each superheat,

as given in Equation (6)

(for 10oC)

(6)(for 5oC)

In normal working conditions, the evaporator surface isalmost covered with condensation water film due to thedehumidification process Therefore, LMED (log meanenthalpy difference) is used to define the overall heat transfercoefficient (Chiou et al., 1994 and Lee et al., 1996b)

(8)

(9)where h aiand h ao represent the enthalpy of inlet and outletair, respectively h ria, h rsa and h roa are saturation wet airenthalpies obtained at the inlet, saturation and outlet refri-gerant temperatures, respectively, and h ro, h rs and h ri meanare the outlet refrigerant, saturation vapor and inlet refri-gerant enthalpy, respectively f e is defined as the ratio ofevaporation capacity at the superheated state to total evapo-ration capacity

Variation of the overall heat transfer coefficient withflow rate of air is shown in Figure 8 Superheats are 5oCand 10oC, ambient temperature is 25oC and relative humi-dity is 50% The overall heat transfer coefficient increaseswith increased air flow rate, and decreased superheat andevaporation pressure Under low flow rates, the superheateffect is small, but the effect is significant for high flowrates

The humidity effect on the overall heat transfer cient for the evaporator is presented in Figure 9, where theambient temperature is 25oC and air volume flow is 420

coeffi-m3/h The overall heat transfer coefficient increases asrelative humidity increases, and the same effects for super-heat and evaporator outlet pressure are seen as in Figure 8

Figure 10 shows the variation of the overall heat transfer

-Figure 6 Condensing capacity and pressure drop for

various refrigerant flow passes

Figure 7 Variation of pressure drop with Reynolds number

for the evaporator Figure 8 Variation of heat transfer coefficient with air flowrate (Ta=25oC, Rh=50%)

PERFORMANCE OF AIR CONDITIONERS WITH A GAS-LIQUID SEPARATION CONDENSER 559

coefficient with ambient temperature, in which flow rate of

air is 420 m3/h and relative humidity is 50% The heat

transfer coefficient increases as ambient temperature

increases

Figure 11 presents the variation of cooling capacity with

air flow rate at 25oC ambient temperature and 50% relativehumidity Cooling capacity increases as the air flowincreases, which is attributed to the increased refrigerantmass flow as a result of higher convection heat transfer athigher air velocity Cooling capacities in superheat 5oC ishigher than those in superheat 10oC

Variation of cooling capacity with relative humidity isshown in Figure 12 The cooling capacity at a relativehumidity of 70% is 23% higher than that at a relativehumidity of 50% High relative humidity causes high airtemperature across the evaporator, and consequently highrefrigerant flow rate

Figure 13 shows the variation of cooling capacity withambient temperature As seen in the figure, cooling capa-city at 30oC ambient temperature is 25% higher than that at

25oC ambient temperature

4 CONCLUSION

Gas-liquid separation type condensers and one-tank ated type evaporators were developed, and heat transferand pressure drop characteristics were investigated experi-

lamin-Figure 9 Variation of heat transfer coefficient with relative

mentally The major results of this study are as follows.

Correlations for pressure drop are derived in terms of

Reynolds number using the least-squares fitting of

experi-mental data

For the condenser, air velocity and mass flow rate of the

refrigerant are main effectors of the overall heat transfer

coefficient; and air temperature and inlet pressure of the

refrigerant slightly affect the overall heat transfer

coeffi-cient

For the condenser, cooling capacity is not significantly

influenced by the refrigerant flow pass, but pressure drop

increases as the superheat region becomes larger

The overall heat transfer coefficient of the evaporator

increases as air flow rate, air temperature and relative

humidity increases, but decreases with increasing

super-heat and outlet pressure

REFERENCES

Ali, A A R., Castro, F., Tinaut, F V and Melrar, M (1995)

Modeling of automotive air conditioning parallel flow

condenser with pressure drop calculations IMechE,

C496/020, 429−434

Chiou, C B., Wang, C C., Chang, Y J and Lu, D C (1994)

Experimental study of heat transfer and flow friction

characteristics of automotive evaporators ASHRAE,

Trans 100, 2, 575−581

Kandlikar, S G (1990) A general correlation for saturated

two-phase flow boiling heat transfer inside horizontal

and vertical tubes J Heat Transfer, 112, 219−228

Kang, J K (1995) Evaporation Heat Transfer and

Pre-ssure Drop of HFC134a and PAG Oil Mixtures in a

Ribbed Flat Channel for Plate/Fin Type Evaporator Ph

D Dissertation National Fisheries University of Pusan

Korea

Kaviany, M (1993) Experimental Study of Pressure Drop

and Heat Transfer in a Plate Evaporator Ford Motor

Co Contract Report, C1-24E

Kurosawa, I and Noguchi, I (1987) Development on ahigh efficiency drawn cup type evaporator core SAE Paper No. 870030

Lee, J H., Jeon, C D., Jeoung, J W and Lee, D K (1996a)

A Study of Pressure Drop and Heat Transfer stics in the Multi-Channel Tube for Automotive Conden- ser Halla Climate Control Co Technical Report HCC-94DE

Characteri-Lee, G H., Jung, J D and Choi, K H (1996b) Systemperformance characteristics of an automotive air condi-tioner with variations of charging conditions Proc SAREK

1996 Winter Annular Conf., 301−306

Masafumi, K (1994) The effective of a cross-section metry on the condensation heat transfer inside multi-passtube Proc WTPF, 2, AFERC, POSTECH, 146−157 Ohara, T and Takahashi, T (1988) High performanceevaporator development SAE Paper No. 880047.Ohara, T., Yamamoto, T and Fujita, H (1990) Heat trans-fer and pressure drop of boiling flow in a cross-ribbedflat channel Int J Heat and Mass Transfer, 17, 555−566

geo-Sugihara, A and Lukas, H G (1990) Performance ofparallel flow condensers in vehicular applications SAE Paper No. 900597

Sahnoun, A and Webb, R L (1992) Prediction of heattransfer and friction for louver fin geometry Trans ASME,

114, 893−900

Webb, R L., Chang, Y J and Wang, C C (1995) Heattransfer and friction correlations for the louver fin geo-metry IMechE, C496/081, 533−541

Yamanaka, Y., Matsuo, H., Tuzuki, K., Tsuboko, T andNishimura, Y (1997) Development of sub-cool system

SAE Paper No. 970110

International Journal of Automotive Technology , Vol 10, No 5, pp 561 − 566 (2009)

561

EXPERIMENTAL STUDY OF THE FLOW CHARACTERISTICS IN AN AUTOMOTIVE HVAC SYSTEM USING A PIV TECHNIQUE

H S JI 1) and S J LEE 2)*

1)Department of Mechanical Engineering, Pohang University of Science and Technology, Gyeongbuk 790-784, Korea

(Current Address; Pusan National University)

2)Department of Mechanical Engineering, Pohang University of Science and Technology, Gyeongbuk 790-784, Korea

(Received 31 October 2008; Revised 2 March 2009)

ABSTRACT− Air flow inside an automotive HVAC module was visualized using a high-resolution PIV technique with varying temperature control modes The PIV (particle image velocimetry) system used for the experiment consisted of a 2- head Nd:YAG laser (125 mJ), a high-resolution CCD camera (2 K × 2 K), optics and a synchronizer A real automotive HVAC module was used as a test model, and some of its casing parts were replaced with transparent windows to capture the flow images of the laser-light-sheet illumination In addition, instant velocity fields were measured for three different temperature control modes by adjusting the temperature baffle Characteristics of the air flow inside the automotive HVAC were then evaluated based on the time-averaged PIV data Results from the experiment showed that flow for the warm mode loses more momentum due to its complicated flow path Thus, the present PIV data can be used to validate numerical prediction and to improve the performance of HVAC modules.

KEY WORDS : HVAC, PIV, Temperature baffle, Bi-level mode, Flow visualization

1 INTRODUCTION

An automotive HVAC (heating ventilating and air

condi-tioning) system is an indispensable device for controlling

temperature, humidity and air ventilation inside the

passen-ger area To improve the performance of an HVAC system

as well as thermal comfort, a deeper understanding of the

characteristics of the air flow inside the HVAC system is

required Recent advancements in computational fluid

dynamics and experimental diagnostic techniques make it

easy to analyze the climatic environment inside a vehicle

In an automotive HVAC module, the temperature baffle

controls the temperature in the passenger area by changing

the direction of air flow

The high-speed air coming from a sirocco-fan does not

easily pass through the HVAC module because of the

temperature control baffles, which curve the flow pathway,

and the heat exchanger Because the air flow inside an

HVAC module has three-dimensional characteristics, it is

therefore not easy to measure it experimentally Hence,

most previous studies for HVAC modules were carried out

using numerical simulations (Aroussi et al., 2001; Bennett

et al., 2002; Kader et al., 2009; Kitada et al., 2001; Shojaee

et al., 2004) Some experimental studies accessed the air

flow inside HVAC modules using point-wise measurement

techniques, such as hot-wire anemometer, 5-hole probe andLDV Hirota et al. (2006) investigated turbulent air mixing

in the T-junction of an HVAC system, while James et al.(2004) investigated a cylindrical HVAC case In addition,Lin (1994) investigated the flow in a simplified HVAC ductsystem due to the technological limitations of conventionalmeasurement techniques

Based on our survey, there is limited quantitative airflow information for HVAC modules Nowadays, PIV(particle image velocimetry) is employed extensively as areliable velocity field measurement technique In this study,

we investigated the characteristics of the air flow inside anHVAC module in detail using a PIV technique

2 EXPERIMENTAL SETUP AND METHODS

The instantaneous velocity fields of air flow in a real

auto-motive HVAC unit were measured using a high-resolution

PIV technique The PIV system used in this study consists

of a CCD camera of 2 K × 2 K pixels resolution,

cylin-drical lens, a dual-head Nd:Yag laser and a delay generator

Figure 1 represents the schematic diagram of the

experi-mental setup Figure 2 shows the experiexperi-mental apparatus

with the laser light sheet illumination The pulse repetition

rate of the dual-head Nd:YAG laser is 15 Hz with an energy

output of 125 mJ per pulse Because the laser pulse has a

short width of about 7 ns, flow images of high-speed air

were captured clearly In order to synchronize the

dual-head Nd:YAG laser and the 2 K×2 K CCD camera, a

delay generator (Stanford DG535) was used The time

interval ∆t between two laser pulses was also controlled

using the delay generator During the time interval ∆t, some

particles moved in and out of the laser light sheet Thus,

there was a need to adjust the thickness of the laser light

sheet pertinently and to shorten the time interval In

addi-tion, the time interval ∆t depended mainly on the maximum

particle displacement in the interrogation window In this

experiment, the laser light sheet was adjusted to about 1

mm thick

Some parts of the HVAC model casing were replaced

with transparent poly-carbonate windows to illuminate thelaser light sheet and to obtain clear flow images Figure 3shows the modified HVAC module with the definition ofcoordinate system used in this PIV experiment The x, yand z axes shown in Figure 3 represent the direction of theexit point of the HVAC module, the downstream flowbeyond the heater exchanger and the exit point of theblower unit, respectively To simulate real operation condi-tions, the air duct system that was connected to the pass-enger area was attached to the end of the HVAC module.This controlled the pressure load acting on the HVACsystem, making the operation conditions almost the same

as those of the real duct system The CCD camera waspositioned perpendicular to the laser light sheet, and avertical plane parallel to the air flow direction was illumin-ated using the laser light sheet This was performed bypassing through the cylindrical lens located in front of themeasurement section

As tracer particles, atomized olive oil droplets weregenerated from two Laskin nozzles The olive oil droplets,with a mean diameter of about 1~2µm, have goodtraceability for the present high-speed flow with largevelocity fluctuations As time goes by, the olive oil dropletsmay deposit on the porous heat exchanger The oil dropletsdeposited on the heat exchanger may affect the mainstream as the flow resistance To prevent obstruction by oildroplets, the optimized experimental conditions werechecked through preliminary experiments From the pre-liminary experiments, optimized raw particle images forobtaining the mean velocity fields were selected as 400image frames In order to obtain accurate instant velocityfield data, the time interval ∆t between the two adjacentparticle images was adjusted adequately using the delaygenerator

Figure 4 shows five measurement sections in the x-yplane for PIV experiments, which measured the two-dimensional whole velocity fields of air flow inside theHVAC module tested in this study The field of view foreach measurement section is 150 × 150 mm2 To evaluatethe whole two-dimensional velocity field of flow inside the

Figure 3 Modified HVAC module and definition of the

coordinate system

Figure 4 Measurement sections in x-y plane

Figure 2 Photograph of experimental apparatus with laser

light sheet illumination

EXPERIMENTAL STUDY OF THE FLOW CHARACTERISTICS IN AN AUTOMOTIVE HVAC SYSTEM 563

HVAC module, velocity field data obtained at five

mea-surement sections were combined together using a

mathe-matical method The mathemathe-matical procedure used for

combining the five measurement sections can be described

as follows

(1) The coordinate for the PIV instantaneous velocity

fields must be adjusted and controlled carefully

(2) To reduce possible errors in the procedure, each

measurement section was overlapped about 20% with the

adjacent sections, as shown in Figure 4

(3) To combine each measurement sections, the

coordi-nate and experimental conditions were substituted through

traditional post-processing procedures

A 50 mm standard lens was attached in front of the CCD

camera to capture the images of air flow inside the HVAC

module

To investigate the effect of different temperature

oper-ation modes on the air flow inside the HVAC system, the

temperature baffle was controlled similar to the real

oper-ation conditions, and the air mix mode was fixed as the

front exhaust mode The three temperature operation modestested in this study were cool-vent mode, warm mode andbi-level mode For comparison purposes, the front exhaustmode was fixed by adopting the same operation conditions.The bi-level mode is usually applied to defrost the frontglass during winter

3 RESULTS AND DISCUSSIONS

At each measurement section, 400 image frames werecaptured consecutively The interrogation window size was

64 × 64 pixels with a 50% overlap From each particleimage pair, the corresponding instant velocity field datawere calculated by applying a cross-correlation PIValgorithm After removing spurious error vectors, the aver-age of the several hundreds of instantaneous velocity fieldswas derived to obtain the time-averaged flow statistics Figure 5 shows the time-averaged mean velocity field inthe horizontal x-z plane in front of the evaporator core Theflow structure in the x-z plane seems to be influenced byFigure 5 Mean velocity fields in the x-z plane

Figure 6 Vorticity distribution in the x-z plane

the shape of the end of the blower unit, the temperature

control mode and the duct shape of HVAC unit located in

front of the evaporator core Figure 5(a) shows the mean

velocity field for the cool-vent mode For the cool-vent

mode, the air passes only the evaporator core and moves

forward to the duct system directly Then the incoming air

flow from the blower unit moves straight along the +x

direction The flow near the exit point of the blower unit

has a high flow speed, but the flows in the regions of

z=0.09 and 0.17 move slowly Moreover, the exit point of

the blower unit has supporting elements to prevent the

deformation of the blower unit The air flow also seems to

be influenced by this supporting structure The maximum

velocity is about 9 m/s at the tip of the blower unit Figure

5(b) shows the mean velocity field for the warm mode

condition The air for the warm mode has to pass both of

the evaporator core and the heater core The maximum

velocity at the tip of the blower is about 6 m/s It is about

30% less than that of the cool-vent mode condition This

indicates that the heat exchanger plays the role of resisting

and reducing the flow momentum in the warm mode

condition

The vorticity contours in the x-z plane in front of the

evaporator core are shown in Figure 6 In the cool-vent

mode, as the flow goes downstream, the vorticity increases

At the locations of x=0.095 and 0.165, the flow has a

counter-rotating vortex structure in the inlet region due to

the high-speed shear flow coming from the blower unit

For the warm mode, the general shape of the vorticity

contours has a similar pattern However, the vorticity for

the warm mode is smaller than that for the cool-vent mode

Figure 7 shows the mean velocity fields in the central

x-y plane For the cool-vent mode, the air flow moves toward

the duct system, directly through the evaporator core as

shown in Figure 7(a) The temperature baffle controls the

air flow moving towards the heater core Due to the

de-flected flow path from the evaporator core to the

connect-ing duct system, stagnation flow is formed in the region

“A” The formation of stagnation region “A” can guide usinto the modification of refrigerant flow path for improvingthe heat exchanging performance The refrigerant flowpath from I to II can be considered better than that from II

to I

In addition, the maximum velocity was about 5.7 m/s infront of the duct system due to the narrow flow path con-necting to the duct Comparing the velocity fields beforeand after the installation of the evaporator core, the flowspeed decreases due to the presence of the evaporator core,which functions as an air flow resistance or a back-pressureloader Thus, the flow of air through the evaporator isclosely related to the performance of the heat exchanger aswell as its flow resistance The air passing through theevaporator core can be observed within the x=0.1 to x=0.24range These flow phenomena shown in the x-y planematch the results in the x-z plane shown in Figure 5

In the warm mode, the air that passes the evaporator coremoves toward the heater cores as shown in Figure 7(b).The path of air flow from the evaporator to the heater core

is nearly perpendicular to the x direction

In addition, the air moves downstream towards theheater core along the flow path After passing through theheater core, the direction of the air turns 180o and enters aconverging flow path

The maximum velocity is about 6.4 m/s at the location(0.24, −0.11) due to the narrowing flow path Stagnationflow is formed in the regions of sudden flow-path vari-ation, such as region B and region B', which are locatednear the heater core From this result, we can see that theperformance of the HVAC module in this condition wasreduced due to the formation of these stagnation regions Inaddition, the formation of stagnation regions B and B’ forthe warm mode can guide us how to design the coolantflow path in order to improve the heat exchanging perfor-mance For the warm mode, to achieve the enhanced heatexchanging performance, the coolant flow path can befrom I to II

Figure 7 Mean velocity distributions in the central x-y plane for two different temperature operation modes

EXPERIMENTAL STUDY OF THE FLOW CHARACTERISTICS IN AN AUTOMOTIVE HVAC SYSTEM 565

To investigate the characteristics of air flow for the

bi-level operation mode, the temperature baffle was

position-ed to divide the air flow passing the evaporator core into

two parts About half of the air passed through the heater

core and moved towards the connecting duct system, while

the other half went directly towards the ventilation duct

This bi-level mode is usually used to defrost the windshield

during winter However, in this study, the front exhaust wasemployed to compare with previous experimental resultsfor other operation modes

Figure 8 shows the mean velocity field for the bi-leveloperation mode The flow from the evaporator core wasseparated into two flow paths around the temperaturebaffle The air flowed directly towards the exhaust duct andFigure 8 Mean velocity distributions in the central x-y plane for the bi-level mode

Figure 9 Schematic diagram for air flow path through HVAC module and refrigerant and coolant path through heatexchangers

accelerated due to the narrowing of effective flow path A

stagnant flow region existed in the upper A of the duct The

maximum velocity of about 5.54 m/s was observed in the

front exhaust duct Even though the stagnation region was

not so widely spread, the performance of the HVAC

module was reduced in this region

The other half of the flowing air moved towards the

heater core The flow moving toward the heater core had a

relatively low speed in the region between the temperature

baffle and HVAC structure The maximum velocity was

about 33% less than the directly ventilating flow This

made the heater core work as a function of back pressure

A large-scale vortex was observed just in front of the heater

core This vortex seemed to be attributed to the steeply

curved flow path due to the flow resistance of the heater

core The performance of the heater exchange can be

influenced by this vortex structure The speed of air

flow-ing through the right part of the heater core was faster than

that in the left vortex formation region From this result, we

can see that if the heated water from an engine block moves

from right to left inside the heater core, heat exchange

performance can be enhanced Due to the flow resistance

of the heater core, the flow behind the heater core has a

relatively low speed In the merging region of flow from

the heater core and from the evaporator core, the maximum

velocity is about 5.43 m/s, which is similar to the cool-vent

mode and warm mode The air passing the heater core

moves toward the duct system through the narrow flow

path Based on these experimental results, the schematic

diagram of the two flow paths are depicted in Figure 9

4 CONCLUSIONS

The entire velocity fields of air flow inside a real HVAC

module were measured using the PIV technique and through

the application of varying temperature operation modes

The results of the present experiment are summarized as

follows

(1) Quantitative information on flow inside a real

auto-motive HVAC module under real operation conditions

was obtained to construct the database and to

under-stand the flow characteristics

(2) The structure of air flow inside the HVAC module was

quite different depending on the temperature operation

mode For the cool-vent mode, the air velocity from the

blower unit moved directly towards the connecting

duct system and through the evaporator core In this

case, the momentum loss of air flow is not high

com-pared to the warm mode On the contrary, for the warm

mode, the air coming from the evaporator core that

passes through the heat core has a maximum flow

speed of about 30% less than that of the cool-vent

mode This means that the performance of a heater core

seemed to be reduced by the formation of the

stagna-tion flow region due to its complicated flow path.(3) The air flow for the bi-level operation mode has theflow characteristics of both the cool-vent mode andwarm mode However, the stagnation flow region forboth operation modes is largely reduced in the bi-levelmode

(4) Based on the detailed flow information around theheater core, we may enhance the performance of heatexchanger by optimizing the flow direction of thecirculating coolant and refrigerant as we mentioned inSection 3 In addition, an inlet/outlet for the heat ex-changer for the coolant and refrigerant can be suggest-

ed from the air flow structure, including the stagnationregions through HVAC module

(5) The present experimental results can be used to validatenumerical predictions and to improve the performance

of automotive HVAC modules in the initial design stage.ACKNOWLEDGEMENT− This work was supported by the Automobile Core Basic Technology Development Project of the Ministry of Commerce, Industry, and Energy of Korea.

REFEREMCES

Aroussi, A., Abdul Ghani, S A A and Rice, E (2001).PIV measurement and numerical simulation of airflowfield in a road vehicle HVAC cowl box SAE Paper No.

2001-01-0294

Bennett, L., Dixon, C W S and Watkins, S (2002) ing and testing of air flow in a HVAC module SAE Paper No. 2002-01-0506

Model-Hirota, M., Asano, H., Nakayama, H., Asano, T andHirayama, S (2006), Three-dimensional structures ofturbulent flow in mixing T-junction JSME Int J Series

Kitada, M., Asano, H., Kanbara, M and Akaike, S (2001).Development of automotive air-conditioning system basicperformance simulator: CFD technique development

JSAE Review, 21, 91−96

Lin, C H (1994) Experimental and computational studies

of flow in a simplified HVAC duct Int J Vehicle Design, 15, 147−165

Shojaee, M H., Tehrani, F P H., Noorpoor, A R andAdili, M R (2004) Analysis of vehicle passenger com-partment HVAC using simulation SAE Paper No. 2004-01-1505

International Journal of Automotive Technology , Vol 10, No 5, pp 567 − 575 (2009)

567

INTEGRATED FUZZY/OPTIMAL VEHICLE DYNAMIC CONTROL

A GOODARZI * and M ALIREZAIE

Automotive Engineering Department, Iran University of Science and Technology, Narmak, Tehran 16765-163, Iran

(Received 28 May 2008; Revised 26 June 2008)

ABSTRACT− There are basically two methods to control yaw moment which is the most efficient way to improve vehicle stability and handling The first method is indirect yaw moment control, which works based on control of the lateral tire force through steering angle control It is mainly known as active steering control (ASC) Nowadays, the most practical approach

to steering control is active front steering (AFS) The other method is direct yaw moment control (DYC), in which an unequal distribution of longitudinal tire forces (mainly braking forces) produces a compensating external yaw moment It is well known that the AFS performance is limited in the non-linear vehicle handling region On the other hand, in spite of a good performance of DYC in both the linear and non-linear vehicle handling regions, continued DYC activation could lead to uncomfortable driving conditions and an increase in the stopping distance in the case of emergency braking It is recommended that DYC be used only in high-g critical maneuvers In this paper, an integrated fuzzy/optimal AFS/DYC controller has been designed The control system includes five individual optimal LQR control strategies; each one, has been designed for a specific driving condition The strategies can cover low, medium, and high lateral acceleration maneuvers on high- µ or low- µ roads A fuzzy blending logic also has been utilized to mange each LQR control strategy contribution level

in the final control action The simulation results show the advantages of the proposed control system over the individual AFS

or DYC controllers

KEY WORDS : Vehicle dynamics control, Integrated chassis systems control, Fuzzy optimal control

1 INTRODUCTION

Recent studies show that active yaw moment control is the

most efficient method for improving vehicle stability and

handling Basically, there are two main methods to control

yaw moment (Van Zanten et al., 1998) The first one is

indirect yaw moment control, which works based on

steer-ing angle control It is simply provided by an actuator that

can add a correction steering angle to the driver's steering

input This technique is referred to as active front steering

(AFS) (Tagawa et al., 1996)

The other method is direct yaw moment control, which

almost works based on differential braking It works by

producing a difference in the longitudinal braking forces on

the two sides of the vehicle to generate the external yaw

moment This technique is referred to as direct yaw moment

control (DYC) (Goodarzi et al., 2003)

Both AFS and DYC are designed as yaw rate and slip

angle controllers DYC is most effective when there are

nearly equal tire-road friction coefficients on both sides of

the vehicle When a µ-split condition exists, the maximum

brake forces on the two sides of the vehicle are different,

and the DYC performance may be limited Another obvious

shortcoming of DYC is the possible reduction of the total

braking force In an emergency situation, a DYC controller

would partially release the brakes on one side of the vehicle,resulting in a longer stopping distance

On the other hand, AFS cannot guarantee satisfactoryperformance in some situations This is due to the inherentnonlinear characteristics of pneumatic tires In normal driv-ing conditions, lateral tire forces are mostly generated withinthe linear region For high lateral acceleration maneuvers,the tire slip angles are usually beyond the saturation limitsand, consequently, the lateral tires forces and correspond-ing yaw moment are not sensitive to the steering input; inthis case, the yaw rate change due to a change in thesteering input would be small (Zeyada et al., 1998) Based on the above discussion, a combined controlstrategy must be used to allow AFS to perform in its effec-tive range while providing the assistance of DYC in thosesituations where it is needed

Several studies have been reported under titles such as

“integrated or global chassis control”, “integrated vehicledynamics control” or “vehicle dynamics management”(Ghoneim et al., 2000; Zeyada et al., 1998; Selby et al.,2001; Trachtler, 2004; Wu et al., 2007) All of thesesstudies approach coordinating the vehicle dynamics func-tions via integrated control of the active chassis systems It

is important, though, to mange the complexity betweenthese systems An appropriate system architecture is requir-

ed, which guarantees a well-defined function for eachsingle system It is important to emphasize that the systems,

*Corresponding author. e-mail: a_goodarzi@iust.ac.ir

even without direct coupling, may be coupled via the vehicle,

and that these indirect couplings may cause unwanted

effects

In this paper, an integrated AFS/DYC control strategy

has been designed The control strategy consists of five

individual optimal strategies, known as Low-g, Mid-g, and

High-g strategies on a high-µ road, and also Low-g and

High-g strategies on a low-µ road Each strategy is suitable

for its corresponding domain To provide continues and

ap-propriate control effort, using fuzzy rules, the five

indivi-dual strategies are blended together On the other word, the

integrated control strategy, intelligently, provides an

appro-priate composition of DYC and AFS control efforts The

simulation results show a good performance of the

integ-rated strategy in different driving conditions

2 STRUCTURE OF THE CONTROLLER

SYSTEM

As is shown in the Figure 1, an Optimal-Fuzzy configuration

has been considered for the proposed integrated AFS/DYC

controller This controller consists of two layers In the

lower layer, there are five individual optimal LQR control

strategies These strategies use the driver steering angle δ,

yaw rate r and lateral velocity ν as inputs and build up the

correction steering angle δ c and DYC external yaw moment

M z as the outputs (Figure 2) The characteristics of each

strategy have been illustrated in Table 1

The final controller’s output is a composition of the five

individual control strategies, as shown below:

(1)where n i, i= 1,2, , 7 are known as participation factors,and their values determine how much each control strategyparticipates in the final controller’s output The participa-tion factors must satisfy the following conditions:

i= 1 7

(2)

n 1, n 2, n 4 and n 6 are calculated by the fuzzy blendingstrategy, which is placed in the higher level of the controlsystem n 3, n 5 and n 7 is achieved as following

The fuzzy blender uses the lateral acceleration a y and vehicleside slip angle β as the inputs to determine which controlstrategy should be triggered and with what contributionstrength The fuzzy strategy also guarantees the continuity

of the final controller’s output signals

3 CONTROLLER DESIGN

3.1 Optimal LQR Control StrategiesAlthough the whole vehicle dynamic behavior and speciallytires behavior are nonlinear, the linear approach is the mostcommon way to express vehicle dynamic behavior.In spite

of the simplicity of the linear models, it is well known that

system

Table 1 LQR strategies’ characteristics

Lateral

Acceleration

Level FrictionRoad ParameterOutput Action

AFS and DYC

Figure 2 Fuzzy/Optimal control process

INTEGRATED FUZZY/OPTIMAL VEHICLE DYNAMIC CONTROL 569

they are only valid in the small slip angles regime, in which

the tires still work in the linear region (Ellis, 1994) In this

study, which is focused on achieving suitable accuracy as

well as benefiting from the simplicity of the linear theories,

a piece-wise linear approach has been used In order to

cover the whole tire working region, the lateral force vs

slip angle curve has been estimated by three lines in the

case of the high friction road’s curve, and two lines in the

case of the low friction road’s curve (Figure 2) The general

form of each linearized piece can be written as:

(4)Considering the above tire model, the common form of the

2DOF vehicle handling model can be modified as follows:

(5)where

Generally, the overall goal of the proposed integrated vehicle

dynamic control system is stated as: the “Minimization of

the difference between the desired and the present motion

of a vehicle”

Note that the desired motion of the vehicles is quite

different from the desired path In order to achieve the

desired motion of the vehicle, its handling and dynamic

behavior must be improved This quality is defined in

accordance with the measurable dynamic parameters of the

vehicles, driver inputs, and some of the safety considerations

Our approach is based on a mathematical presentation ofthe desire motion

We shall define a general performance function I, known

as the lateral vehicle dynamic index:

(6)where r d, v d are the desired yaw and lateral velocityresponses, which can be analytically calculated based onthe reference vehicle model (Ghoneim et al., 2000):

(7)Considering the steady state response, the desired valuescan be calculated as:

(8)The function I is defined in such a manner that its mini-mization corresponds to the desired motion achievement

By controlling the external yaw moment, M z and thecorrection steering angle δ c the performance index of (6)must be minimized The two performance criteria beingconsidered are the lateral velocity and the yaw rate, whichare both used simultaneously

The first term in equation (6) is used to reinforce the yawrate following characteristics of the vehicles The secondterm in (6) is the lateral velocity, with its analogousquantity of side slip angle, is another essential parameter inthe vehicle dynamics Researchers indicate that, in a safemaneuver on a dry road performed by a normal driver, thisangle is limited to about 2 degrees and, for professionaldrivers, it would be limited to about 4 degrees (Van Zanten

et al., 1998) Therefore, in order to maintain the safety ofthe vehicles, we must avoid any increase in the values ofside slip angle or lateral velocity Using the both termsguarantees an improvement of the safety and stability ofthe vehicle with good handling

The final point that must be considered is the physicallimitations of the external yaw moment M z and the correc-tion steering angle δ c These are affected by some limita-tions that result from the steering and braking systems’characteristics and the road friction Consequently, the con-trol process obtained without considering any limitations isnot practical In order to consider the above mentionedlimitations, we shall include the external yaw moment andcorrection steering angle into the performance function

By adjusting the weighting factors w 1 to w 3 in (6), therelative importance between each term is not only deter-mined, but also satisfies all the physical limitations

By expressing the performance index (6) in the belowmatrix form, an infinite-time-horizon LQR problem is de-fined (Kirk, 1970):

2bCr – aC f

I zz u

- 2b2Cr + a 2 C f

I zz u -–

E=

2C f

m

2aC f