Fuel Reactivity Controlled Compression Ignition: A

Fuel Reactivity Controlled Compression Ignition: A Pathway to High-Efficiency, Clean Combustion Rolf D. Reitz, Reed M. Hanson, Sage L. Kokjohn, Derek A. Splitter

January 28, 2011 Acknowledgments DERC Member Companies DOE/Sandia National Labs

1

SAE Government Industry meeting

Outline • RCCI combustion background • LD and HD Engine and experimental setups • Results – Comparison of light- and heavy-duty engines at 9 bar IMEP (single cylinder engine experiments) – Methods to reduce heat transfer losses in the lightduty engine (CFD modeling)

• Fuel effects • Current research directions • Conclusions

2


RCCI Background

• • • •

With appropriate control HCCI and PCCI can provide a fuel efficiency advantage over mixing/diffusion limited strategies – Improved start- and end-of-combustion  Minimize compression work and maximize expansion work; Low temperature combustion  minimize emissions Fuels: Dec et al. (2010-01-1086)  TE=48% mid-size engine w/ gasoline boosted HCCI Manente et al. (2010-01-1471)  TE=57% HD engine w/ RON=80 gasoline (RI, soot?) Combustion phasing control still remains a challenge The addition of a second fuel allows significant control over the auto-ignition characteristics of the charge 0.1 900

100

750

80

600

60

450

40

300

20

150

0 -20 -15 -10

-5

0

5

10

15

20

25

0 30

[J/  ]

RCCI

Ignition Delay [sec]

Pressure [MPa]

120

1050

Conventional Diesel

Heat Release Rate

140

n-heptane (diesel fuel) 50-50 blend of gasoline and diesel fuel iso-octane (gasoline)

0.01

1000

1E-3

100

1E-4

10

1E-5 600

IC Engine Regime 700

800

900

1000

1100

1 1200

Temperature [K] Diesel SOI [ATDC]

Crank [ATDC]

3


Ignition Delay [CA @ 1300 rev/min]

•

Multiple injection strategy - RCCI combustion* Port injected gasoline

Direct injected diesel

Injection Signal

Gasoline

Squish Conditioning

Ignition Source

Diesel

-80 to -50

-45 to -30

Crank Angle (deg. ATDC) *ENGINE COMBUSTION CONTROL VIA FUEL REACTIVITY STRATIFICATION WARF US Patent Application: 09820924-P100054, 2/2010

4


Heat Release Analysis • Heat release occurs in 3 stages (SAE 2010-01-0345) – Cool flame reactions resulting from n-heptane (diesel) injection – First portion of energy release occurs where both n-heptane and iso-octane (gasoline) are mixed – Final portion of the energy release occurs where primarily isooctane (gasoline) is located

• Changing fuel delivery ratio changes relative magnitudes of each stage

200

200

Cool Flame

PRF Burn

Primarly n-heptane

n-heptane + entrained iso-octane

o

Iso-octane Burn

80 C & 56% iC8H18 o

150 AHRR [J/o]

AHRR [J/o]

150

109 C & 68% iC8H18

Primarly iso-octane

100

50

91% iC8H18 80% iC8H18

o

140 C & 80% iC8H18 o

170 C & 91% iC8H18

68% iC8H18

56% iC8H18

100

50

0

0 -20

-10

0 o Crank [ ATDC]

10

5

20

-20

-10 0 o Crank [ ATDC]

10


20 5

Experimental Setups Engine

Heavy Duty

Light Duty

CAT SCOTE

GM 1.9 L

Displ. (L/cyl)

2.44

0.4774

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

0.7

2.2

IVC (°ATDC)

-85 and -143

-132

EVO (ATDC)

130

112

Engine

Swirl ratio

Injector type Nozzle holes Hole size (µm)

Diesel Gasoline/ E85 Heavy-Duty Engine

Light-Duty Engine

Common rail 6

8

250

128

6


Heavy-Duty Engine Results

Experiment = Solid

120

Simulation = Dash

Soot [g/kW-hr]

140

100 80

Gross Ind. Efficiency [%]

60 40 20 0 800

4.6 bar

700 600 500 400 300

5.9 bar 9.3 bar 11.6 bar 14.6 bar

200 100 0 -20 -15 -10 -5

0

5

10 15 20 25 30

Crank [ATDC]

Kokjohn et al. IJER 2011 Hanson et al. SAE 2010-01-0864 Splitter et al. THIESEL 2010

0.3

2010 EPA HD Limit

Experiment Simulation

0.2 0.1 0.0

Ringing Int. 2 [MW/m ]

AHRR [J/deg]

Pressure [bar]

RCCI gives near zero NOx and soot and peak gross thermal efficiency of 56% (59% with diesel/E85 –Thiesel, 2010)

NOx [g/kW-hr]

RCCI – high efficiency, low emissions

0.02

2010 EPA HD Limit

0.01 0.00 56 54 52 50 48 4 2 0 4

6

8

10

12

14

IMEPg [bar]

7


16

Light- and heavy-duty RCCI: 9 bar IMEP • Heavy-duty and light-duty engines compared over gasoline-to-diesel fuel ratio sweep at 9 bar IMEP Engine Engine

Heavy Duty

Light Duty

CAT

GM 1.9 L

IMEP (bar)

9

Speed (rev/min)

Total Fuel Mass (mg)

1300

1900

94

20.2

EGR (%)

41

Premixed gasoline

82 to 89

81 to 84

Diesel SOI 1 (°ATDC)

-58

-56

Diesel SOI 2 (°ATDC)

-37

-35

Diesel Inj. Pressure (bar)

800

500

Intake Pressure (bar)

1.74

1.86

32

39

1.75

0.46

Intake Runner Temp. (°C)

Air flow rate (kg/min) Port fuel

Gasoline

DI fuel

Diesel Fuel

8


Light- and heavy-duty RCCI: 9 bar IMEP NOx [g/kW-hr]

0.3

2010 EPA HD Limit 81%

0.1

Low NOx and soot is achieved for both HD and LD engines Ringing intensity is easily controlled by combustion phasing (via adjustments to the gasoline-diesel ratio) with only minimal effect on efficiency Both engines achieve high efficiency; however, HD engine shows 5 to 7% higher thermal efficiency

Light-duty

82%

•

84% 82%

0.0 2010 EPA HD Limit

0.02

Soot [g/kW-hr]

•

Heavy-duty

0.2

• 0.01

56 54 52 50 48

140

2.8

Heavy-Duty: 89% Gasoline Experiment

100

6

2.4

Simulation

2.0

Light-Duty: 83% Gasoline Experiment

80

1.6

Simulation

60

1.2

40

0.8

2

20

0.4

0

0

0.0

3 bar/deg.

4

-1

0

1

2

3

4

5

6

CA50 (o ATDC) IMEPg [bar]

7

8

-30

-20

-10

0

10

20

Crank [ATDC]

9


30

Heat Release Rate [1/ms]

120

Pressure [bar]

Ringing Int. 2 [MW/m ]

Gross Ind. Efficiency [%]

0.00

Light- and heavy-duty RCCI: 9 bar IMEP 60

56

Fuel Energy [%]

Heavy-duty Light-duty

50

50 40

31 31

30

20 11

15

10 2

4

0 ITE gross

• • •

Heat Transfer Comb. Loss

Combustion efficiency is ~2% lower for the light duty engine  Crevice geometry must be improved Heat transfer losses account for remaining differences in gross indicated efficiency Light-duty engine has higher heat transfer losses due to: – – –

10

Exhaust

Higher swirl ratio Larger surface-to-volume ratio Lower mean piston speed


Light- and heavy-duty RCCI: 9 bar IMEP • To isolate effects of combustion process and heat transfer, adiabatic operation was explored using CFD modeling (KIVA-CHEMKIN) • Thermal efficiencies (rather than gross indicated efficiencies) are compared to isolate influence of heat transfer • Thermal efficiency of HD and LD engines are nearly identical when heatHeavy transfer removed - Dutyis Engine

100

0.20

80

0.16

60

0.12

40

0.08

20

140 120

0.04

Light - Duty Engine

0 -20 -15 -10

Pressure [MPa]

0.24

-5

0

5

10

Experiment Crank [ATDC] Sim - With HT Sim - Adiabatic

100

15

20

25

Light Duty

0.00 30

0.28 0.24 0.20

80

0.16

60

0.12

40

0.08

20

0.04

0 -20 -15 -10 -5

0

5

10

Crank [ATDC]

15

20

25

closed

0.00 30

11

Thermal Efficiency [%] (-132 to -112 deg. ATDC)

Pressure [bar]

Sim - Adiabatic

[1/  ]

Sim - With HT 120

thermal 

0.28

Heat Release Rate

Heavy Duty

Experiment

Heat Release Rate [1/]

140

60

W132 to 112 Fuel Energy*combustion

Heavy-duty

57.2 57.7

Light-duty

55 51.0 50

47.5

45 40 With Heat Transfer

Adiabatic

*Work evaluated for both engines from -132° to 112° ATDC


Methods to reduce heat transfer losses Factor Swirl Speed (rev/min) Geometry Heavy-duty

High Level

Low Level

2.2 2239 Scaled Light-duty

0.7 1900 Base Light-duty

Light-duty

Scaled LD

•

Reduction in heat loss (%) 6 -2 4.7

23 full factorial DOE –

Combustion phasing was held constant at 2˚ ATDC by adjusting the PRF number of the premixed fuel

• Reducing swirl ratio reduces heat transfer losses by 6% • Reducing surface-to-volume ratio reduces heat transfer losses by 4.7% • Interaction of swirl ratio and piston bowl geometry is positive and additive

12


Methods to reduce heat transfer losses Heavy- LightLD duty duty Improved 0.01 0.04 0.03 0.01 0.01 0.01 2.7 3.7 4.8 54 47 53 10.9 13.8 11.6

ISNOx (g/kW-hr) ISsoot (g/kW-hr) Ringing Int. (MW/m2) Gross Ind. Eff. (%) Heat Loss (%)

13

100

Fuel Energy [%]

• With improved combustion chamber geometry and reduced swirl ratio, light-duty results are very similar to heavy-duty results • Thermal efficiency is increased by 2.2 % of the fuel energy • Remaining increase in gross indicated efficiency results from improvements in combustion efficiency

80 60 40 20

Comb. Loss (%) Heat Transfer (%) Exhaust (%) ITE gross (%)

0

Heavy-duty

Light-duty

LD Improved

* Bars sum to greater than 100 because gross indicated efficiency (GIE) is presented


Fuel effects - GDI engine? Additized gasoline* gasoline/1.75 % DTBP gasoline E-85/diesel gasoline/diesel

Same Peak HTHR Location 9.6 bar IMEPg 140

1.4

0.020

2010 HD limit

0.015

1.0

0.005

0.03 0.02 0.01 0.00 -20

-15

-10

-5

Crank Angle (CA ATDC)

0

1.75% DTBP 90% port fuel 43% EGR

60

E-85 78% port fuel 0% EGR

Gasoline/Diesel 89% port fuel 43% EGR

40 20

0.8 0.6 0.4

14

0.000

10 6 2 4.0 3.5 3.0 2.5 2.0 1.5 1.0 0.5 0.0

0.2

0

-25

0.010

0.0

-20

-15

-10

-5

0

5

10

15

20

25

Crank Angle ( CA ATDC)

* Splitter et al. 2010-01-2167

14

0.3

NOx (g/kw-hr)

• Engine does not run without DTBP/EHN • DI gasoline plus 1.75% additive same performance as DI diesel  DTBP dosing ~0.2% of total fuel rate • NOx, soot below EPA 2010 • ISFC 145 g/kW-hr, 56% TE

0.2

2010 HD limit 0.1 0.0 4

6

8

10

12

14

16

IMEPg (bar)


18

COV (%)

80

1.2

0.04

PRR (bar/ CA)

100

0.05

AHRR (kJ/ CA)

Pressure (bar)

120

AHRR (kJ/ CA)

NTC Behavior

PM (g/kw-hr)

DI gasoline w/ cetane improver DTBP: di-tert-butyl peroxide

Current Research – Sandia National Labs • Chemiluminescence and soot luminosity images -10

-5

-3

0

2

5

10

15

UNIBUS

-15

Crank = -3 deg. ATDC

Crank = 2 deg. ATDC

Crank = 4 deg. ATDC

Crank = 6 deg. ATDC

Crank = 9 deg. ATDC

Crank = 12 deg. ATDC







Crank = 2 deg. ATDC

Crank = 4 deg. ATDC

Crank = 7 deg. ATDC












300

UNIBUS RCCI Flame Propagation

250 AHRR [J/o]

Flame Propagation

RCCI


200 150 100 50 0 -30

15

-20

-10 0 10 20 Crank [deg. After Peak AHRR]

30


Conclusions • RCCI combustion has been demonstrated in a light-duty engine using single cylinder engine experiments – Results show near zero NOx and soot emissions and gross indicated efficiency greater than 48% over wide range of loads – Controlled energy release (control over combustion phasing)

• Port fuel injection of gasoline (cost effective), direct injection of diesel or additized gasoline (low injection pressure). Diesel or GDI (w/spark plug) operation retained. • Comparisons made between light- and heavy-duty RCCI combustion at similar operating conditions. • Thermal efficiency in light-duty engine is 5 to 7% lower due primarily to increased heat transfer losses. • CFD modeling was used to investigate methods to reduce heat transfer losses in light-duty engine – With similar swirl, mean piston speed, and combustion chamber geometry, light-duty RCCI combustion shows nearly identical results to heavy-duty engine

• RCCI technology provides practical low-cost pathway to >20% improved fuel efficiency (lower CO2), while meeting emissions mandates in-cylinder

16


Impact of Fuel Efficiency on US Oil Consumption • In 2008, US economy paid an average of $28.5 billion/month to buy foreign oil - 10,000 gal/s ~ 25% of world total • US Petroleum consumption: 20.7 Million Barrels of Oil per Day* 65% used in transportation = 13.5 MBOD Potential Truck and Automotive fuel usage reduction by Dual Fuel: 4.2 MBOD Diesel: 45%  53% = improvement of 18% = 0.6 million barrels saved 9.3 MBOD Gasoline SI: 30%  53% = improvement of 77% = 4.1 million barrels saved Total saved = 4.7 MBOD = 34% of US transportation oil (23% of total US petroleum used ~ $1 Billion saved / 2 days)

• Could reduce transportation oil consumption by 1/3 = US imports from Persian Gulf

- while surpassing 2010 emissions regulations • Consistent with US DOE/EERE FreedomCar & 21st Century Truck fuel efficiency goals: 50% increase in LD, 25% increase in HD

*

http://www.eia.doe.gov/

17


RCCI with light load operation (2.2 bar) Port fuel = Gasoline DI fuel = Gasoline + 3.5% 2-EHN

800 [rev/min]

0.005 0.000

0.3 0.2 190

0.1

185 180 175 170

EPA 2010 HD

0.010

CO (g/kwh)

0.4

800 [rev/min]

35

25

15

15

10 5 0

NOX (g/kwh)

 gross (-)

0.50 0.48 0.46 0.44 0.42 0.40

1.5 1.0

EPA 2010 HD

0.5 0.0

-1

0

1

2

CA50 ( ATDC)

SAE 2011-01-0361

18

-1

0

1

CA50 ( ATDC)


2

HC (g/kwh)

14 12 10 8 6 4 2 0

HTHR Duration (Deg) PM (g/kwh)

1300 [rev/min] 0.015

ISFC (g/kwh)

PFI Percent [-]

1300 [rev/min]

Port fuel = Gasoline DI fuel = Gasoline + 3.5% 2-EHN

Combustion Phasing Control At a fixed combustion phasing of 2° ATDC, iso-octane delivery ratio (i.e., global PRF number) is linearly dependent on intake temperature – Operation is premixed enough to exhibit HCCI-like tradeoff between PRF number and intake temperature – SENKIN single zone simulations show very similar trend in required PRF to achieve desired combustion phasing

• •

Fuel delivery ratio provides CA50 control over a wide range of intake temperatures With an accurate mechanism, single zone modeling can be used as a starting point for selection of iC8H18 delivery ratio

90 85

GDI%            0.4  o  Intake Temp ( C)     

80 75 70

RCCI Experiments CR SOI = -50° ATDC

65 60 55

80

90

95 90 85

   PRF  o 0.48  Temp (  Intake C)    

75

SENKIN Single Zone ERC PRFv1 Adiabatic Tivc = Tivc_exp + 3 K

70 65 60 55

19

100 110 120 130 140 150 160 170 Intake Temperature [oC]

80

PRF

•

Delivery Ratio [% iso-octane]

95

80

90

100 110 120 130 140 150 160 170 Estimated Intake Temperature [oC] 19
