Reactivity Controlled Compression Ignition (RCCI) … Engine Research Center, 2016 Reactivity Controlled Compression Ignition (RCCI) for high-efficiency clean IC engines Prof. Rolf

1 Engine Research Center, 2016

Reactivity Controlled Compression Ignition (RCCI)

for high-efficiency clean IC engines

Prof. Rolf D. Reitz

Engine Research Center,

University of Wisconsin-Madison

Aurel Stodola Lecture

Department of Mechanical and Process Engineering of ETH Zurich

Wednesday, November 9th, 2016

Acknowledgements:

Industry Partners:

Direct-injection Engine Research Consortium members,

Caterpillar, Ford, General Motors.

Sandia, Argonne, Oak Ridge National Labs,

ARO, DOE, NASA, ONR, Princeton CEFRC.

- ERC faculty, staff and students

UW-Madison

Founded in 1848 (land-grant institution). 42,600 students; 2,000 facultyRanked No. 2 US public university

~ $1 billion/year in research funding

Engine Research CenterLargest academic research center focusing

on IC engines in the US

~ $4-5M annual research budget

~ 8 Faculty, 55 graduate students,

10-15 research staff, post doc/visiting scholars

Mechanical Engineering35 faculty, 7 instructional staff

- 543 undergraduates

- 170 graduate students

~$10 million in research annually


Engine Research Center

3

Myers BormanUyehara

Founded in 1946 ARO Center of Excellence; DOD DOE; Consortia/Industry

2-color soot radiation,

droplet experiments

0-D engine simulation,

heat release analysis

1985 - 2003

First to demonstrate HCCI in a 4-stroke engine

First to demonstrate benefits of multiple injections for power and emissions

Pioneered Multi-Dimensional modeling for concept evaluation

Advanced spray/combustion diagnostics for fuel-air mixing and emissions characterization

In-cylinder gaseous species measurement, high resolution spray analysis

High-fidelity engine codes, detailed chemistry for emission prediction, spray & turbulence modeling

Control strategies, HCCI/RCCI

Combustion system optimization

Aftertreatment experiments and modeling

2003 +

Engine Research Center, 2016

Emeritus faculty

Mario Trujillo Sage KokjohnScott Sanders

ERC faculty

Outline


• Motivation for investigating Internal Combustion (IC)

engine efficiency

• Requirements for high-efficiency, clean combustion

• Engine combustion modeling and optimization

background

• Reactivity Controlled Compression Ignition (RCCI):

- pathway to high-efficiency clean combustion using in-cylinder

blending of fuels with different auto-ignition characteristics

• Performance of RCCI combustion

• Limits and practical applications of RCCI

• New concepts

• Conclusions and future research directions

Why research IC engine efficiency?


Internal combustion (IC) engines are the work horses of transportation and

power generation

70% of all crude oil consumed is used to fuel IC engines

Improvements in engine efficiency can have a major impact on fossil fuel

consumption and green house gas (GHG) emissions on a global scale

IC engines are expected to be the dominant (>90%) prime mover for

transportation applications well into the future1,2,3

Many open questions:4

- what combustion process should future IC engines utilize?

- what fuels are best suited for high efficiency combustion?

- how can renewable and alternative fuel sources be utilized most

effectively?

1Quadrennial Technology Review, DOE 20112Review of the Research Program of the FreedomCAR and Fuel Partnership: 3rd

Report, NRC 2010 3Energy Information Agency, Annual Energy Outlook 2012, June 2012. 4Reitz, R.D., “Directions in Internal Combustion Engine Research,” CNF, 160, 2013.


Boyd T (1950) Pathfinding in Fuels and

Engines. SAE 500175, 4(2) 182-195.

Fuel ignitability affects engine efficiency - limits compression ratio (CR).

Early Spark Ignition (SI) engines were plagued by “spark knock”, CR ~ 4:1.

Cylinder pressure measurements by Midgley and Kettering at DELCO/GM

showed different fuels had different knock tendency

e.g., kerosene worse than gasoline

Volatility differences were thought to be the explanation.

Guided by the “Mayflower,” they added a red dye (iodine) to kerosene

and knock tendency was greatly reduced!

Unfortunately, tests with other red dyes did

not inhibit knock, disproving the theory.

But, finding powerful antiknock additives was

a major serendipitous discovery!

Mayflower – Trailing Arbutus Jane

in early spring

Lessons from history (1910-20) – “the Mayflower”1/11 CR


Research after WW-I was motivated by national security - Improved fuel efficiency with higher CRs made possible the first

non-stop airplane flight from New York to San Diego in the 1920’s.

GM and US Army studied hundreds of additives and found aromatic

amines to be effective knock suppressors.

1920 experimental GM car driven on gasoline with toluidine with CR ~7:1 - 40% better fuel consumption than 4:1.

Engine exhaust plagued by unpleasant odors

Lessons from history (1920-30) – the Amines and TEL

Reitz, Front. Mech. Eng. 1:1, 2015

Much research was devoted to find acceptable additives, - finally leading to tetraethyl lead (TEL)

But, TEL caused solid deposits, damaged exhaust valves and spark plugs.

Scavenger additives with bromine and chlorine corrected the problem.- Partnership with Ethyl-Dow and DuPont to extract compounds from sea water

- 10 tons of sea water needed to provide 1 lb of bromine!

- “the goat”!

WW-II aviation engines used iso-heptane (triptane: 2,2,3-trimethyl butane) - allowed CR as high as 16:1.


Lead poisoning was an early concern - In 1926 US Surgeon General determined that TEL poses no health hazards.

- Use of lead in automotive fuels has been called “The mistake of the 20th century”

1950: Dr. Arie Haagen-Smit - cause of smog in LA to be HC/NO - Cars were the largest source of UHC/NOx

1950: Eugene Houdry - developed catalytic converter for auto exhaust. - But, lead was found to poison catalytic converters.

20 years later: US EPA announces gas stations must offer "unleaded" gasoline,

- Based accumulated evidence of negative effects of lead on human health.

- Leaded gasoline was still tolerated in certain applications (e.g., aircraft),

but was permanently banned in the US in 1996, in Europe since 2000

Lessons from history (1930-70) –TEL and the future

Reitz, Front. Mech. Eng. 1:1, 2015

World Wars & national security played a major role to define automotive fuels.

Today’s engines and their fuels would not have been developed without

close collaboration between engine OEMs, energy and chemical companies!

A consequence of collaboration between “big” engine and “big” oil is that

transformative changes in transportation systems will not occur easily.

A new concept engine must be able to use available fuels,

A new fuel must run in existing engines.

Maximizing engine efficiency

• Fuel energy is wasted due to:

– Incomplete combustion (i.e.,

combustible material flowing out

the exhaust)

– Heat transfer losses to the

coolant, oil, and air

– Unrecovered exhaust energy

– Pumping losses

– Friction losses

• Goal is to maximize Brake

Thermal Efficiency

• Diesel engines have more

than 10% higher thermal

efficiency than Spark-Ignited

engines


CAFE

On July 29,

2011

President

Obama

announced an

agreement

with 13 large

automakers to

increase fuel

economy to

54.5 miles per

gallon for cars

and light-duty

trucks by

model year

2025

Sankey diagram

http://en.wikipedia.org/wiki/Model_year

Diesel emissions regulations

US Environmental

Protection AgencyEuropean Union

After-treatment: considerable expense and complexity – VW scandal!

In addition to increasing fuel efficiency, stringent mandates must be met.



Advanced diesel combustion

Conventional diesel Low temperature H/PCCI

• High EGR to reduce temperature,

early injection to improve mixing

• Low NOx and soot emissions

• Start of combustion controlled by

by fuel chemistry

- limits engine load (less power)

• Combustion timing and heat

release controlled by injection

• High NOx and soot emissions

Sandia Cummins

N14 optical engine

HCCI


Advanced diesel combustion

Conventional diesel Low temperature H/PCCI

• High EGR to reduce temperature,

early injection to improve mixing

• Low NOx and soot emissions

• Start of combustion controlled by

by fuel chemistry

- limits engine load (less power)

• Combustion timing and heat

release controlled by injection

• High NOx and soot emissions

Sandia Cummins

N14 optical engine

HCCI

Engine combustion optimization via CFD modeling

Engine combustion physics occurs over an

immense range of time and length scales

Unique modeling challenge that includes

virtually all aspects of mechanical engineering

- Turbulence

- Combustion

- Two-phase flow

Chemistry of real fuels involves 100’s of

species and 1000’s of reactions

Bore

diameter

Turbulence-

chemistry

interactions

10-2 10-6 10-9

Soot

Inception

10-3 10-6 10-9

Engine

Cycle

Fluid flow Chemistry

Length Scale [m]

Time Scale [s]

Ideally all scales would be resolved considering all species and reactions with

Direct Numerical Simulation (DNS) and detailed chemistry

single engine cycle would take decades

Engineering applications must model:

Small scales of turbulence (RANS or LES)

Turbulent combustion (average species production)

Chemical kinetics (reduced, but representative chemistry)

In current practical CFD models all scales of turbulence are typically

modeled (Reynolds Averaged Navier-Stokes (RANS)) and combustion is

modeled using direct-integration of a reduced chemical kinetics mechanism



UW-ERC Multidimensional CFD modelsSubmodel Los Alamos UW-Updated References

intake flow assumed initial flow compute intake flow SAE 2011-01-0821

heat transfer law-of-the-wall compressible, unsteady E&F 2012

turbulence standard k- RNG k- /LES SAE 2011-01-0829

nozzle flow none cavitation modeling SAE 2011-01-0386

atomization Taylor Analogy surface-wave-growth SAE 960633

Kelvin Hemholtz SAE 980131

Rayleigh Taylor CST 171, 1998

drop breakup Taylor Analogy Rayleigh Taylor AAS 6, 1996

drop drag rigid sphere drop distortion SAE 960861

wall impinge none rebound-slide model SAE 880107

wall film/splash SAE 982584

collision/coalesce O’Rourke shattering collisions AAS 19, 2009

vaporization single component multicomponent fuels IJMF 35, 2009

low pressure high pressure IJHMT 45, 2002

ignition Arrhenius reduced chemistry CNF 158, 2011

combustion Arrhenius CTC/GAMUT SAE 2004-01-0102

Sparse matrix/cluster SAE 2012-01-1974

NOx Zeldo’vich GRI reduced SAE 2008-01-2413

soot none 2-Step formation/oxidation IJTS 48, 2009

PAH/Multi-step phenomen. CST 182, 2010

R/D

L/D

Breakup lengthL Ca

12/ f(T )

Blob

injection

model

Spray model improvements at the ERC

Discrete drop

model

=0et

r=B0

KH Model

RT Model

Kelvin-Helmoltz

Rayleigh Taylor

Linearized instability analysis

KH-RT

Spray Models

Nozzle flow/cavitation

Jet atomization

Drop breakup

Drop collision/coalescence

Drop distortion and drag

Multi-component fuel evaporation

Spray-wall impingement



Surrogate fuels - 18 component model

alkanes

aromatics

cycloalkanes

PAH

corrected

DMC vaporization model: Diesel, Jet-A, JP-8

Fuel property and chemistry surrogates

Cyclo alkanes

n-alkane

n-alkane

iso-alkanes

Aromatics

cyclohexane

decalin cyclohexane

n-dodecane

n-heptane

n-octadecane

n-tetradecane

naphthalene

mcymene

tetralin

n-pentylbenzene

n-heptylbenzene

toluene

heptamethyl nonane

tetramethyl hexane

iso-octane

Physical property

surrogates

Chemistry

surrogates

Chemical class grouping – MultiChem skeletal mechanism

25 species

51 reactions

857 species

3586 reactions

LLNL detailed

mechanismERC reduced

mechanism*

* Ra and Reitz, Combustion & Flame 2011

Mu

ltiC

he

mM

ech

an

ism

10

0 s

pe

cie

s, 3

48

re

actio

ns


Primary Reference Fuels

6 bar IMEP

0 10 20 30 40 50 600

20

40

60

80

100

250

180 g/kW-hr

MISFIRE

190

g/kW-hr

PR

F [-]

EGR Rate [%]

10

bar/deg.

5.6

bar/deg.

Combustion optimization - fuel and EGR selection

HCCI simulations used to choose

optimal EGR rate and PRF

(isooctane/n-heptane) blend

at 6, 9, and 11 bar IMEP, 1300 rpm

Predicted contours agree well

HCCI experiments

As load is increased the minimum

ISFC cannot be achieved with

neat diesel or neat gasoline

0 10 20 30 40 50 600

20

40

60

80

100

190 g/kW-hr

180 g/kW-hr

PR

F [-]

EGR Rate [%]

10 bar/deg.

0 10 20 30 40 50 600

10

20

30

40

50

60

70

80

90

100

24 bar/deg

28 bar/deg

160

MISFIRE

190 g/kW-hr

170 g/kW-hr

180 g/kW-hr

PR

F

EGR Rate [%]

170

190

210

230

240

200

Net ISFC

[g/kW-hr]

9 bar IMEP

0 10 20 30 40 50 600

20

40

60

80

100

16

bar/deg.

230

220

180

MISFIRE

210

190

200

PR

F [-]

EGR Rate [%]

11 bar IMEP

EGR Rate [%]

PR

F

Kokjohn, SAE 2009-01-2647

Gasoline-diesel

“cocktail”

Gasolin

eD

iesel


KIVA CFD plus Genetic Algorithm optimization used to choose injection parameters

- Gasoline (I-C8H18)

- Diesel (n-C7H14)

Fuel reactivity and EGR from

HCCI investigation (9 bar IMEP)

Global PRF = 65

EGR rate = 50%

80% gasoline/20% diesel

- SOI1 ~ -60°ATDC

- SOI2 ~ -33°ATDC

- 60% of diesel fuel

in first injection

-25 -20 -15 -10 -5 0 5 10 15 201E-6

1E-5

1E-4

1E-3

0.01

oh

HRR

Temp

ch2o

conc7h16

c14h30

ic8h18

ch2o

oh

co

(J/Deg)

(K)

Crank [ATDC]

Mass F

raction [-]

ic8h18

0

200

400

600

800

1000

1200

1400

1600

HR

R [J/

] and A

vera

ge T

em

pera

ture

[K

]

Gasoline

Diesel

Controlled combustion

timing and duration

Charge preparation optimization


Kokjohn & Reitz ICLASS 2009

Premixed and Direct Injected fuel blending

Direct injected diesel

Port injected gasoline

-80 to -50 -45 to -30

Crank Angle (deg. ATDC)

Inje

cti

on

Sig

nal Squish

Conditioning

Ignition

Source

Gasoline

Optimized Reactivity Controlled Compression Ignition (RCCI)

Optimized fuel blending in-cylinder

Gasoline Diesel

Diesel



Replace Diesel Exhaust Fluid (DEF) with widely available fuel

Heavy- and light-duty ERC experimental engines

Engine Heavy Duty Light Duty

Engine CAT SCOTE GM 1.9 L

Displ. (L/cyl) 2.44 0.477

Bore (cm) 13.72 8.2

Stroke (cm) 16.51 9.04

Squish (cm) 0.157 0.133

CR 16.1:1 15.2:1

Swirl ratio 0.7 2.2

IVC (°ATDC) -85 and -143 -132

EVO(°ATDC) 130 112

Injector type Common rail

Nozzle holes 6 8

Hole size (µm) 250 128

LDHD

Engine size scaling

Staples, SAE

2009-01-1124


IMEP (bar) 9

Speed (rpm) 1300

EGR (%) 43

Equivalence ratio (-) 0.5

Intake Temp. (°C) 32

Intake pressure (bar) 1.74

Gasoline (% mass) 76 82 89

Diesel inject press. (bar) 800

SOI1 (°ATDC) -58

SOI2 (°ATDC) -37

Fract. diesel in 1st pulse 0.62

IVC (ºBTDC)/Comp ratio 143/16

Computer modeling predictions confirmed

Combustion timing and Pressure Rise Rate control with diesel/gasoline ratio

Effect of gasoline percentage

Hanson, SAE 2010-01-0864

-30 -20 -10 0 10 20 300

2

4

6

8

10

12

14 Experiment

Simulation

Crank [ATDC]

Pre

ssu

re [M

Pa

]

0

200

400

600

800

1000

1200

1400

89%

Gasoline

76%

82%

89%

App

are

nt H

ea

t R

ele

ase

Ra

te [J/]

Neat

Gasoline

Neat

Diesel Fuel

Dual-fuel can be used to extend load limits of either pure diesel or gasoline

Experimental validation - HD Caterpillar SCOTE


4 6 8 10 12 14 16

4548515457

0.00

0.01

0.02

0.030.0

0.1

0.2

0.3

Heavy-duty RCCI (gas/gas+3.5% 2-EHN, 1300 RPM)Heavy-duty RCCI (E-85/Diesel, 1300 RPM)Heavy-duty RCCI (gas/diesel 1300 RPM)

Gro

ss

Ind

.E

ffic

ien

cy

Gross IMEP [bar]

So

ot

[g/k

W-h

r]

HD Target (~2010 Levels)

NO

x[g

/kW

-hr]

HD Target (~2010 Levels)

RCCI – high efficiency, low emissions, fuel flexibility

Splitter, SAE 2010-01-2167; Hanson, SAE 2011-01-0361

Indicated efficiency of 58±1%

achieved with E85/diesel

Emissions met in-cylinder,

without need for after-treatment

Considerable fuel flexibility,

including ‘single’ fuel operation


Diesel can be replaced with

<0.5% total cetane improver

(2-EHN/DTBP) in gasoline

- less additive than SCR DEF

Dual-fuel RCCI combustion – controlled HCCI

Heat release occurs in 3 stages

Cool flame reactions from diesel (n-heptane) injection

First energy release where both fuels are mixed

Final energy release where lower reactivity fuel is located

Changing fuel ratios changes relative magnitudes of stages

Fueling ratio provides “next cycle” CA50 transient control

24

-20 -10 0 10 200

50

100

150

200

Crank [oATDC]

AH

RR

[J/o

]

Primarly iso-octane

n-heptane+ entrained iso-octane

Iso-octane BurnPRF BurnCool Flame

Primarly n-heptane

80 90 100 110 120 130 140 150 160 17055

60

65

70

75

80

85

90

95

Intake Temperature [oC]

De

live

ry R

atio

[%

iso

-octa

ne

]

CA50=2 ˚ ATDC

RCCI

SOI = -50 ATDC

RCCI

Reitz et al. US Patent 8,616,177 - 2013


HCCI

Model validation via RCCI optical experiments

RCCI experiments in Sandia HD

optical engine

LED illumination through side windows

to visualize sprays

Images recorded through both piston-

crown and upper window

Crank-angle-resolved high-

temperature chemiluminescence

with high-speed CMOS camera

Short-wave pass filter to reject long-

wavelength (green through IR) soot

luminosity

Engine Cummins N-14

Bore x stroke 13.97 x 15.24 cm

Displacement 2.34 L

Geometric compression ratio 10.75

GDI

Iso-octane

100 bar

7x150 micron

Common-rail

n-heptane

600 bar

8x140 micron

Inc. Ang. 152°Kokjohn, SAE 2012-01-0375


26 Engine Research Center, 2016Bowl window Squish (upper) window

Load: 4.2 bar IMEP GDI SOI: -240°ATDC

Speed: 1200 rpm CR SOI: -57°/-37° ATDC

Intake Temperature: 90° C Equivalence ratio: 0.42

Intake Pressure: 1.1 bar abs. Iso-octane mass %: 64

Kokjohn, ILASS-2011

RCCI combustion luminosity imaging


Computational Optimization of RCCI

SAE 2011-24-0015


• Sharp end-of-combustion

• Peak temperatures reduced

significantly compared to

conventional diesel combustion

• Difference between peak and

average temperatures very small

for RCCI combustion

– Well mixed combustion event-20 -15 -10 -5 0 5 10 15 20 25 300

20

40

60

80

100

120

140

+4

TDC

-4

-8-12

-16

RCCI

Diesel

Crank [ATDC]

Pre

ssu

re [

bar]

Diesel SOI

Retard

-20

0

300

600

900

1200

1500

Hea

t R

ele

as

e R

ate

[J

/]

-20 -15 -10 -5 0 5 10 15 20 25 30600

900

1200

1500

1800

2100

2400

2700

+4

TDC

-4

-8

-12

-16

RCCI

Diesel

Crank [ATDC]

Pe

ak

Te

mp

era

ture

[K

]

-20

-20 -15 -10 -5 0 5 10 15 20 25 30600

900

1200

1500

1800

2100

2400

2700

+4TDC-4-8

-12-16

RCCI

Diesel

Crank [ATDC]

Avera

ge T

em

pera

ture

[K

]

-20

Comparison of conv. diesel and RCCI - KIVA CFD

9 bar IMEP,

1300 rpm,

41% EGR


Temperature contours near CA50, Crank = 6.1 deg ATDC

Comparison of conv. diesel and RCCI combustion

• High temperature in conventional diesel next to piston bowl surface

• Highest temperature for RCCI in center of chamber (adiabatic core)

• Region near liner has similar temperatures

– heat transfer differences are at piston bowl surface


RCCI

(800 bar)

High-EGR

Diesel

(1800 bar)

Ringing (MW/m2) 3.8 2.8

Max PRR (bar/deg) 10.3 8.9

Gross Indicate TE (%) 54.3 48.5

Combustion Loss (%) 1.3 0.7

ISNOx (g/kW-hr) 0.006 0.97

ISsoot (g/kW-hr) 0.01 0.1

48.557.9

19.2 10.9

31.641.5 33.4 38.1

61.954.3

0

10

20

30

40

50

60

70

80

90

100

With HT Adiabatic With HT Adiabatic

High-EGR Diesel RCCI

Fu

el E

nerg

y [

%]

Comb. LossExhaustHeat LossGIE

Comparison of conv. diesel and RCCI - KIVA CFD

RCCI:

- Similar pressure rise rates

- Significantly lower NOx and soot

- 16% higher thermal efficiency

- Reduced heat losses (~50%)

- Improved end-of-combustion phasing

SOI -12o

-10 -5 0 5 10 15 20 25

39

42

45

48

51

54

57

RCCI

High EGR Diesel Pinj 800 bar

High EGR Diesel Pinj 1800 bar

RCCI Combustion

SOI -20ATDC

CA50 [ATDC]

Gro

ss In

dica

ted

Eff

icie

ncy

[%]

7% of

fuel energy

Pancake 18.7:1 piston design

~1.2 less surface area

GT-Power heat transfer HX tuned to

match data

- 14.9:1 piston required HX = 0.4

- 18.7:1 required HX = 0.3

- without oil cooling, HX = 0.2

GTE

(%)

IMEPg

(bar)

NTE

(%)

IMEPn

(bar)

Experiment 59.1 6.82 55.0 6.27

Model, HX =0

100% comb. η62.4 7.12 58.5 6.85

Model, HX =0

100% comb.η,

0% EGR

63.4 7.23 61.0 6.95

94% of maximum theoretical cycle efficiency achieved !

-40 -30 -20 -10 0 10 20 30 40-15

0

15

30

45

60

75

90

105

120

135

150

E85 / 3% EHN+91 PON RCCI

43C intake, 42% EGR,

6.3 bar IMEPn

EXP, Squirter off, 43% EGR, Oil Matrix Point 83

GTPower, HX=0, 100% comb. , 43% EGR

GTPower, HX=0, 100% comb. , 0% EGR

Pre

ss

ure

(b

ar)

Crank Angle (CA ATDC)

0

150

300

450

600

750

AH

RR

(J

/ C

A)

Splitter et al. “RCCI Engine Operation Towards 60% Thermal Efficiency”, SAE 2013-01-0279


Ultra high efficiency, dual fuel RCCI combustion

DI: 3%EHN+91ON

PFI: E85

TIVC = 43 C

EGR = 42%

16.1

14.918.7


GM 1.9L Engine Specifications

Multi-cylinder RCCI - transient operation, open/closed loop control

Engine Type EURO IV Diesel

Bore 82 mm

Stroke 90.4 mm

Displacement 1.9 liters

Cylinder

Configuration

Inline 4

4 valves per cylinder

Swirl Ratio Variable (2.2-5.6)

Compression

Ratio 17.5

EGR SystemHybrid High/Low

Pressure, Cooled

ECU (OEM) Bosch EDC16

ECU (new) Drivven

Common Rail

Injectors

Bosch CRIP2-MI

148° Included Angle

7 holes,

440 flow number.

Port Fuel

Injectors

Delphi

2.27 g/s steady flow

400 kPa fuel pressure

UW RCCI Hybrid Vehicle

SAE 2015-01-0837

Highway Fuel Economy Testing of an RCCI Series Hybrid VehicleReed Hanson, Shawn Spannbauer, Christopher Gross, Rolf D. Reitz, University of Wisconsin; Scott Curran, John Storey, Shean Huff, ORNL

http://www.sae.org/servlets/productDetail?PROD_TYP=PAPER&PROD_CD=2015-01-0837


. UDDS/FTP cycle

RCCI offers diesel-like or

better BTE across speed-load range

ORNL simulations indicate RCCI

offers >20% fuel economy c.f.

2009 PFI engines

Wagner, Aramco

Workshop, 2014

ERC/ORNL Collaboration - RCCI - vehicle operation

RCCI Emissions


ORNL experiments indicate that > 95% of the PM in RCCI is condensed organic hydrocarbons.

RCCI low particle number – by 2 orders of magnitude


Collaboration with Oak Ridge – GM 1.9L Multi-cylinder engine

SAE 2010-01-2266


Operating Condition Low-

Load

Mid-Load High-Load

Gross IMEP [bar] 4 9 11 13.5 16 23

Engine Speed [rpm] 800 1300 1370 1460 1550 1800

Intake Press. [bar abs.] 1.00 1.45 1.94 2.16 2.37 3.00

Intake Temp. [°C] 60 60 60 60 60 60

Caterpillar 3401E SCOTE

Displacement [L] 2.44

Bore x Stroke [mm] 137.2 x 165.1

Con. Rod Length [mm] 261.6

Compression Ratio 16.1:1

Swirl Ratio 0.7

IVC [deg ATDC] -143

EVO [deg ATDC] 130

Common Rail Diesel Fuel Injector

Number of Holes 6

Hole Diameter [μm] 250

Included Spray Angle 145o

Design Parameter Minimum Maximum

Premixed Methane [%] 0% 100%

DI Diesel SOI 1 [deg ATDC] -100 -50

DI Diesel SOI 2 [deg ATDC] -40 20

Diesel Fraction in First Inj. [%] 0% 100%

Diesel Injection Pressure [bar] 300 1500

EGR [%] 0% 60%

Natural gas/diesel RCCI

ERC KIVA PRF kinetics

NSGA-II MOGA

32 Citizens per generation

~9500 Cells @ BDC

UW Condor -

Convergence after ~40 genrtns


Nieman, 2012RCCI Fuel flexibility – Alternative fuels

Double vs. triple injection

Results 2 Inj. Optimum 3 Inj. Optimum

Soot [g/kW-hr] 0.004 0.004

NOx [g/kW-hr] 0.24 0.10

CO [g/kW-hr] 10.8 7.3

UHC [g/kW-hr] 10.5 3.8

ηgross [%] 45.1% 47.1%

4 bar IMEP 23 bar IMEP

Results 2 Inj. Optimum 3 Inj. Optimum

Soot [g/kW-hr] 0.079 0.014

NOx [g/kW-hr] 0.08 0.17

CO [g/kW-hr] 6.0 1.7

UHC [g/kW-hr] 9.4 3.3

ηgross [%] 44.1% 46.5%

44.1% 42.4%

7.9% 5.6%

46.5%43.0%

8.5% 2.0%0%

5%

10%

15%

20%

25%

30%

35%

40%

45%

50%

Gross Work Exhaust Loss Heat Transfer Combustion Loss

% o

f F

uel

En

erg

y I

n2 Inj. Optimum3 Inj. Optimum

45.1%

31.5%

17.1%

6.3%

47.1%

31.9%

18.7%

2.4%0%

5%

10%

15%

20%

25%

30%

35%

40%

45%

50%

Gross Work Exhaust Loss Heat Transfer Combustion Loss

% o

f F

uel

En

erg

y I

n

2 Inj. Optimum3 Inj. Optimum


Nieman MS 2012RCCI Fuel flexibility – Alternative fuels

Advanced fueling strategies – DDFS


2015: Wissink, M.L., PhD - Direct Injection for Dual Fuel Stratification

(DDFS): Improving the Control of Heat Release in Advanced IC Engine Combustion Strategies

Wissink & Reitz SAE 2015-01-0856

RCCI

PPC

PPC

DDFS

DDFS ignition timing controlled by RCCI.

Load expansion achieved by late gasoline injection.

Combustion stability considerably improved.

Advanced fueling strategies – DDFS efficiency


~14% reduction ~12% reduction

~1% increase

~30% increase

~13% reduction

DDFS provides high efficiency, lower noise/COV, lower heat

loss/increased exhaust loss – reducing turbo requirements

Conclusions and future research directions

Advanced combustion strategies (e.g., GCI, RCCI, DDFS and variants) offer practical low-cost pathways to >15% improved internal combustion engine fuel efficiency (lower CO2)

Made possible by advances in fuel injectors and computer control

RCCI GTEs in the 58-60% range achieved within ~94% of theoretical cycle.

Inconvenience of two fuels already accepted by diesel industry (diesel/DEF)

RCCI is cost effective and offers fuel flexibility: - low cost port-injected less reactive fuel (e.g., gasoline, E85, “wet” EtOH, C/LNG) with

optimized* low pressure DI of more-reactive fuel (e.g., diesel/additized gas) - reduced after-treatment needed - meet NOx and PM emission mandates in-cylinder- diesel or SI (w/spark plug) operation can be retained (e.g., mixed mode, limp home).

Improved transient control:- proportions of low and high reactivity fuels can be changed dynamically, with same/next-

cycle combustion feedback control

Direct injection of both fuels allows more control of heat release:- reduced noise, reduced cyclic variability, no efficiency penalty, move waste heat to exhaust

Future directions:- transient engine feedback control, load extension (e.g., via: multiple injection, CR, VVA),- optimized pistons – reduced crevice volumes, insulated pistons.- optimized boost, EGR, charge-air cooling, alternative fuels……

.. and vehicle testing!


SAE 2013-01-0279

* WARF Pat. 9,376,955B2

2025 and beyond….

Voyage to new concepts in engine combustion

California ARB:

90% reduction in NOx emissions by 2031 (0.02 g/bhp·hr)

80% reduction in GHG emissions (below 1990 levels) by 2050

50% petroleum reduction target by 2030 (renewable fuels),

and continued reductions in air toxics & diesel PM (PN 6×1011 1/km).

In a chronically leaking boat, energy devoted to changing vessels is likely

to be more productive than energy devoted to patching leaks.

Warren Buffett, Investor


Backup

DOE Secretary Steven Chu announced Jan. 11th, 2012 at Detroit Economic Club:

“Research we’ve supported by the University of Wisconsin and Sandia National

Labs has shown that low-temperature combustion could considerably improve

engine efficiency and increase the fuel economy of light-duty vehicles by more

than 50%”

Documents

Reactivity Controlled Compression Ignition (RCCI) … Engine Research Center, 2016 Reactivity Controlled Compression Ignition (RCCI) for high-efficiency clean IC engines Prof. Rolf