VECOM Low Temperature Combustion

Low Temperature Combustion 

Stanislav V. Bohac 

Associate Research Scientist, University of Michigan 

Visiting Scientist, Czech Technical University in Prague 

VECOM Workshop, Prague, March 14-15, 2012

My background: 

Degrees 

B.S. & M.S. in Mechanical Engineering, University of Illinois, USA, 1993 &1995 

Ph.D. in Mechanical Engineering, University of Michigan, USA, 2002 

Positions in Industry 

AVL LIST, Graz, Austria, 1995-1997 

AVL Powertrain, Plymouth, MI, USA, 1998-2002 

Current Positions 

Associate Research Scientist, University of Michigan, USA 

Visiting Scientist, VECOM, Czech Technical University in Prague, Czech Republic 


My research interests: 

Using thermodynamics, combustion analysis, and emissions characterization to 

study and develop advanced combustion and emissions reduction systems for 

the transportation industry. 


Part 1: Gasoline 



Outline 

• Why consider low temperature combustion (LTC) for gasoline engines 

• Tools used to study LTC 

• Improving efficiency with LTC 

– Naturally aspirated LTC 

– Boosted LTC 

• Ignition and combustion constraints of LTC and the need for three 

combustion modes: HCCI, SACI, SI 

• Combustion phasing 

• Ringing and misfire limits 

• Emissions 

• Summary 

• Future challenges 

VECOM Workshop 5 

Prague, March 14-15 2012

Why consider low temperature combustion (LTC) 

for gasoline engines 



Normal (High) Temperature Combustion 

Good control of combustion 

Combustion phasing is set with spark timing 

Good emissions characteristics with TWC: 

U.S., 2012, EPA Tier 2, Bin 5 U.S., 1960, unregulated 

NMOG = 0.075 g/mi THC = 10.6 g/mi 99% ↓ 

CO = 3.4 g/mi CO = 84.0 g/mi 96% ↓ 

NO x = 0.05 g/mi NO x = 0.41 g/mi 88% ↓ 

PM = 0.01 g/mi 

↓ 

Poor efficiency as compared to diesel engines 

2010 GM 2.0L LNF Gasoline: 323 g/kWh @ 2000 rpm, 3 bar BMEP 18% ↑ 

2009 VW 2.0L TDI Diesel: 265 g/kWh @ 2000 rpm, 3 bar BMEP 




for gasoline engines 

To improve engine efficiency while maintaining 

good emissions characteristics. 



Tools for studying LTC 



Tools for Studying LTC 

1. 1-D Cycle Simulation: GT Power 

2. Single-cylinder LTC research engine 

3. Multi-cylinder LTC research engine 

1-D Cycle Simulation 

Single-Cylinder LTC Research Engine 

Multi-Cylinder LTC Research Engine 

VECOM Workshop 10 Prague, March 14-15 2012

Improving efficiency with LTC 


Carnot Cycle Efficiency 

Carnot Cycle 

• The theoretically most efficient cycle for converting thermal energy into work. 

• Heat taken from hot reservoir (T h ) and rejected to cold reservoir (T c ) 

• 1-2: isothermal heat addition 

• 2-3: isentropic expansion 

• 3-4: isothermal heat rejection 

• 4-1: isentropic compression 

T 

1 

T 

c 

h 

• Efficiency improves with higher T h and lower T c 


Carnot Cycle Efficiency 

For a gasoline engine, one might be tempted to use: 

T h = 2400 K (typical peak cylinder temperature) 

T c = 363 K (engine coolant temperature of 90°C) 

and then claim that maximum attainable efficiency is: 

Tc 

363 

1 

1 

85% 

T 2400 

h 

It appears that LTC (lower T h ) would reduce efficiency. 

This logic is a mistake! 

‣ IC engines do not follow the Carnot Cycle. 

‣ The theoretically most efficient cycle for IC engines is the Otto Cycle. 


Otto Cycle Efficiency 

Otto Cycle: The theoretically most efficient cycle for an IC engine. 

• 1-2: isentropic compression 

• 2-3: constant volume heat addition 

• 3-4: isentropic expansion 

• 4-1: constant volume heat removal 

1 

1 

CR 

 

( 1) 

CR = compression ratio = V BDC /V TDC 

γ = C p /C v = ratio of specific heats 

Otto Cycle efficiency is a function of CR and γ. 


Otto Cycle Efficiency 

1 

1 

CR 

 

( 1) 

To increase efficiency we have two options: 

1) Increase compression ratio, CR 

2) Increase ratio of specific heats, γ 


Otto Efficiency (%) 

Increasing Compression Ratio 

 

1 

1 

CR 

 

( 1) 

V 

CR 

V 

BDC 

TDC 

65 

60 

γ=1.29 

55 

50 

45 

40 

35 

6 8 10 12 14 16 18 20 22 24 

Compression Ratio (-) 


Increasing Compression Ratio – Real Behaviour 

 

calc 

 

1 

 

CR 

1 

1 

 

ideal Otto cycle, 

• γ=1.29 

Wi 

i 

 

m Q 

Wb 

b 

 

m Q 

f 

f 

LHV 

LHV 

work on piston 

• heat loss 

• comb. duration 

• comb. efficiency 

• pumping 

work at flywheel 

• friction 

* 

Caris, D. F., Nelson, E. E., “A New Look at High Compression Engines,” SAE Technical Paper 590015. 


Increasing Compression Ratio 

Increasing CR beyond 12-14 provides little efficiency gain: 

• Increased heat losses (higher surface area to volume ratio) 

• Decreased combustion efficiency (fuel trapped in crevices, quenching) 

Furthermore, increased CR: 

• Requires higher octane fuel (expensive) 

• Requires stronger engine structure (high peak cylinder pressure) 

Conclusion: Increasing CR beyond current range for naturally aspirated engines 

(10.5-11.5) does not have the potential for large improvements in efficiency. 



 

1 

1 

CR 

 

Increasing Ratio of Specific Heats 

( 1) 

 

C 

C 

p 

v 

65 

60 

CR=11 

55 

50 

45 

40 

35 

1.20 1.22 1.24 1.26 1.28 1.30 1.32 1.34 1.36 1.38 1.40 

C p /C v (-) 


1 

1 

CR 

 


( 1) 

 

C 

C 

p 

v 

equilibrium products 

 

( F / A) 

( F / A) 

s 

 

1 

 

Gamma and engine efficiency increase with: 

• Lower temperature 

• Leaner (lower phi) 




equilibrium products 

65 

60 

CR=11 

55 

50 

45 

40 

35 

1.20 1.22 1.24 1.26 1.28 1.30 1.32 1.34 1.36 1.38 1.40 

C p /C v (-) 

EGR dilution increases gamma by decreasing temperature 

Air dilution increases gamma by decreasing temperature and changing composition. 

Raising gamma from 1.22 to 1.28 increases ideal Otto efficiency by 19%, and this does not even include 

pumping benefits at low load. 

Conclusion: Significant efficiency benefits can be attained by increasing gamma. 


Motivation for Low Temperature Combustion 

Objective: Increase engine efficiency 

↓ 

Increase efficiency by increasing gamma 

↓ 

Increase gamma using low temperature, lean combustion 


Real Engine Effects 

Real engines do not have constant gamma. 

Real engine efficiency is reduced by: 

• Heat Loss 

• Burn duration loss 

• Combustion efficiency 

• Pumping loss 

• Friction loss 

A GT-Power 1-D cycle simulation was set up to account for these effects. 

Simulation results adapted from: 

• Lavoie, G. A., Ortiz-Soto, E., Babajimopoulos, A., Martz, J. B., Assanis, D. N., “Thermodynamic Sweet Spot Under Highly Dilute and 

Boosted Gasoline Engine Conditions”, submitted to International Journal of Engine Research, March 2012. 

• Martz, J., “Towards Efficient IC Engines”, presented at ACCESS Combustion Meeting, Ann Arbor MI, March 2012. 

• Assanis, D., Lavoie, G., Ortiz-Soto, E., “Thermodynamic Sweet Spot Under Highly Dilute and Boosted Gasoline Engine Conditions”, 

SAE 2011 High Efficiency Engines Symposium, Detroit MI, April 10-11, 2011. 


Engine Model 

GT Power used to simulate a 4 cylinder, 2.5L, current design engine: 

Bore/Stroke 

CR 12 

Intake Valves (2) 

IVO (0.1 mm lift) 

IVC (0.1 mm lift) 

Exhaust Valves (2) 

EVO (0.1 mm lift) 

EVC (0.1 mm lift) 

Combustion 

90mm / 100 mm (4 cylinders, 2.5L) 

32.4 mm diameter / 10.8 mm lift 

12°BTDC gas exchange 

224°ATDC gas exchange 

26.1 mm diameter / 10.8 mm lift 

135°ATDC firing 

371°ATDC firing 

Wiebe profile 

Heat transfer Woschni 1 

Friction Chen-Flynn 2 

1) Woschni, G., 1967, “Equation for the Instantaneous Heat Transfer Coefficient in the Internal Combustion Engine”, SAE Paper 670931. 

2) Chen, S. K., Flynn, P. F., 1965, “Development of a Single Cylinder Compression Ignition Research Engine.” SAE Paper 650733. 


Engine Model – Operating Conditions 

GT Power model of 4 cylinder, 2.5L, current design engine: 

Engine speed 

Phi 0.2 - 1.2 

EGR 0-80% 

T in 60°C 

S P 

8 m/ 

s 

(2400 rpm) 

P in 

1 - 3 bar 

Turbocharger Eff. (overall) 40, 50, 60% 

10-90% burn 25°CA, fixed for all cases 

CA50 

10°ATDC, fixed for all cases 


Effect of Phi on Naturally Aspirated Engine Efficiency 

70 

eta_all_v_phi_effect of h-b-pump-fric 

(%) 

60 

50 

40 

30 

Gr. 

Gr. 

Gr. 

Net 

Brake 

TIMING LOSS 

PUMPING 

HEAT LOSS 

FRICTION 

BURN 

DUR. 

2° 

25° 

25° 

25° 

20 

0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 

 

For Otto cycle, leaner is always better. 

For brake efficiency, best efficiency is = 0.5 - 1.0 

• Heat loss and friction reduce efficiency at lower phi. 

• Lower gamma and incomplete combustion reduce efficiency at higher phi. 


Effect of Phi on Naturally Aspirated Engine Efficiency 

 

( F / A) 

( F / A) 

s 

 

' 

 

F /( A R) 

( F / A) 

s 

For all cases, heat transfer and friction decrease efficiency at low BMEP. 

EIVC reduces pumping losses. 

EGR dilution reduces pumping and increases gamma. 

Air dilution further increases gamma. 


Effect of Phi on Boosted Engine Efficiency 

60 

etag etab_v_phi_PIN eq PEX_var Pin 

50 

GROSS 

P IN 

= P EX 

3 bar 

2 bar 

1 bar 

(%) 

40 

30 

BRAKE 

BOOST 

20 

0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 

Boosting slightly improves gross efficiency as heat loss becomes proportionally smaller 

(Nu=Re m ). 

Boosting greatly improves brake efficiency as friction becomes proportionally smaller 

(friction is mainly a function a speed, not engine load). 

 

Boosting moves peak brake efficiency to leaner phi (best = 0.3 - 0.6 for P in =P exh =3 bar). 

AIR 


Effect of Phi on Boosted Engine Efficiency with TC Eff.=50% 

_brake (%) 

50 

40 

30 

etab_v_bmep_AIR_TC50 

AIR DILUTION T/C EFF = 50% 

0.3 

0.4 

’ 

0.5 

0.6 

1.2 1.5 2.0 2.5 3.0 

P_in (bar) 

1.0 

Throttled SI 

( = 1) 

20 

0 10 20 30 40 

BMEP (bar) 

Boosted SI 

( = 1) 

For all cases, heat transfer and friction decrease efficiency at low BMEP. 

Boosting helps at =1 (friction and heat loss are less significant). 

For a given boost, leaner is better (higher gamma) until friction and heat loss dominate. 

For best efficiency, boost and phi should increase together. 

1.0 


Optimum Phi and Boost with TC Eff.=50% 

BMEP (bar) Phi (-) P in (bar) 

5 0.35 1.2 

10 0.45 2.0 

15 0.50 2.5 

20 0.55 3.0 

25 0.60 3.0 

30 0.65 3.0 

35 0.95 3.0 

At low loads, phi and boost should increase together with BMEP. 

After maximum possible boost is reached, phi should increase alone. 


Ignition and combustion constraints of LTC, and the 

need for three combustion modes: HCCI, SACI, SI 


Ignition and Combustion with LTC 

The analysis so far has assumed that: 

• Homogeneous charge 

• The charge is ignitable 

• Combustion is sustainable 

• Combustion proceeds at a reasonable pace 

But ignition and combustion are not guaranteed… 


Ignition and Combustion with LTC 

Consider three ranges of Phi: 

1) = 0.65 – 1.0 → SI (spark ignition) 

• As becomes leaner approaches 0.65, ignition occurs but 

combustion becomes unstable through expansion. 

2) = 0.20 – 0.45 → HCCI (homogeneous charge compression ignition) or CAI 

• Preheat charge using intake heating or negative valve overlap (NVO) 

‣ NVO closes exhaust valve early and opens intake valve late (recompression) 

• Compression initiates combustion 

• < 0.2 is not practical with typical CR 

• > 0.45 causes excessive combustion rate (noise, engine damage) 

3) = 0.45 – 0.65 → SACI (spark assisted compression ignition) 

• Ignite charge with spark; this consumes part of charge at a moderate 

rate 

• Initial combustion compresses remaining charge (like a 2 nd piston), 

and compression ignition consumes the rest. 

• Attain complete combustion with a tolerable combustion rate. 


Combustion Phasing 

Combustion phasing is set by charge temperature at IVC. 

Two methods: 

1) Intake heating 

• Useful in a research engine because have independent 

control over temperature and composition 

2) Negative valve overlap (NVO) 

• Useful in a production engine because it is fast acting and 

does not require a powerful intake heater. 


Ignition and Combustion with Naturally Aspirated LTC 

BRAKE 

(%) 

50 

40 

30 

20 

HCCI 

0.3 

Load control_etab_v_bmep_NA 

0.4 

ADV. 

COMB 

’ 

SI 

0.6 1.0 

AIR DILUTION 

EGR DIL. ( = 1) 

EIVC (SI; = 1) 

THROTTLED (SI; = 1) 

10 

0 2 4 6 8 10 12 

BMEP (bar) 


Combustion phasing 


Combustion Phasing 

The simulations assumed that CA 50 is maintained at 10°ATDC. 

Is this optimum 

Is this possible 


Combustion Phasing – SI Combustion 

What is optimum combustion phasing 

gross (%) 

60 

55 

50 

45 

40 

etag_timing sweeps 2-10-20-30-40 

10-90 BURN = 

2 ° 

10 ° 

CR = 12; = 0.6 

20 ° 

30 ° 40 ° 

ADIABATIC 

WITH H.T. 

35 

-10 TDC 10 20 

With no heat transfer, optimum CA 50 = TDC. 

With heat transfer, optimum CA 50 = 10°ATDC. 

Burn duration has minimal effect on optimum CA 50 . 

Burn duration reduces efficiency for burn durations greater than 20°. 


Ringing and misfire limits 


Ringing and Misfire Limits of HCCI 

HCCI combustion phasing is not as flexible as for SI combustion. 

• Advanced combustion causes excessive rate of heat release. 

• Undesirable combustion noise 

• Increased thermal and mechanical loading of engine. 

• Retarded combustion causes unreliable combustion. 

• Cycle-to-cycle changes in IMEP. 

• Misfires 


HCCI Ringing Limit 

HCCI and SACI can cause excessive rate of heat release (ringing). 

* 

Vavra, J., Bohac, S. V., Manofsky, L., Lavoie, G., Assanis, D., 2011, “Knock in Various Combustion Modes in a Gasoline-Fueled Automotive 

Engine,” Proceedings of the ASME 2011 Internal Combustion Engine Division Fall Technical Conference, ICEF2011-60124. 





Three commonly used parameters to quantify the ringing limit: 

1) Ringing intensity 1 

Based on the assumption that that engine structure excitation is 

proportional to the acoustic intensity of the cylinder pressure 

waves. 

Apply a 5 kHz high pass filter to each cylinder pressure trace: 

RI 

 

P 

P 

 

2 

1 

max RT 

, 2 

2 

max 

W 

m 

 

 

P max 

Keep RI < 5 MW/m 2 

time 

1 

Andrea M. M., Cheng W. K., Kenney T, Yang J., On HCCI Engine Knock, SAE Paper No. 2007-01-1858, 2007. 



2) Low pass ringing index 1 

Based on the assumption that ringing is initiated by and is 

proportional to the maximum rate of low freq. pressure rise. 

Apply a 5 kHz low pass filter to each cylinder pressure trace 

RI LP 

 

 

 

1 

2 

For typical passenger car engine geometry, 

β=0.05 

dP 

dt 

 

 

 

This approach is useful for engine simulations 

that do not capture high frequency oscillations. 

Keep RI < 5 MW/m 2 

 

 

 

W 

 

m 

max 

RT 

, 

2 

Pmax 

2 

-20 -10 0 10 20 30 40 

1 

Eng J. A., “Characterization of Pressure Wave Oscillation in HCCI Combustion,” SAE Paper No. 2002-01-2859, 2002. 

VECOM Workshop 44 Prague, March 14-15 2012 

 

 

 

60 

55 

50 

45 

40 

35 

30 

25 

P 

t 

 

 

 

dP 

dt 

 

 

 

max 

P max


3) Combustion Noise 1 

Combustion Noise Meter (e.g., AVL 450S) is commonly used. 

Perform Fourier transform of cylinder pressure trace. 

Apply “U filter” to account for engine structure attenuation. 

Apply “A filter” to account for reception of human ear. 

U filter characteristics 

A filter characteristics 

0 

5 

Magnitude [dB] 

-10 

-20 

-30 

Magnitude [dB] 

0 

-5 

-10 

-15 

-40 

0.1 1 10 

freq [kHz] 

-20 

0.1 1 10 

freq [kHz] 

Combustion noise (dB) = RMS of resulting signal 

Keep Combustion Noise < 90 dB 

1 

Russell M.F., Haworth R., Combustion Noise from High Speed Direct Injection Diesel Engines, SAE Paper No. 850973, 1985. 


Comparison of HCCI Ringing Parameters 

Experimental setup to compare HCCI ringing metrics: 

• Single cylinder research engine 

• Hydraulic valve actuation for NVOcontrolled 

HCCI 

• Bore / stroke = 86.0 mm / 94.6 mm 

• CR = 12.5 

• Direct injection, side-mounted injector 

• 91 RON gasoline 


HCCI Ringing Limits 

Comparison of three commonly used ringing metrics using an HCCI 

single cylinder research engine 1 . 

HCCI 

SACI 

SI 

Different methods do not always show the same trends for different 

combustion modes. 

1 

Vavra, J., Bohac, S. V., Manofsky, L., Lavoie, G., Assanis, D., 2011, “Knock in Various Combustion Modes in a Gasoline-Fueled Automotive 

Engine,” Proceedings of the ASME 2011 Internal Combustion Engine Division Fall Technical Conference, ICEF2011-60124. 


HCCI Misfire Limit 

COV of IMEP increases as CA 50 is retarded 1 . 

COV limit at later phasing with lower ON fuels 

3 

NH40 

2.5 

RD387 

COV of IMEP g 

2 

1.5 

1 

Iso-Octane 

0.5 

0 

0 2 4 6 8 10 12 14 

CA50 [deg ATDC] 

1 

Hagen, L., Bohac, S., Lavoie, G., Assanis, D., “Fuel Effects on Heat Release and Maximum Load with HCCI in the UM FFVA Engine,” 

presented at Sandia National Lab, Livermore CA, February 2011. 


IMEP g [kPa] 

Combining HCCI Ringing and Misfire Limits 

Both ringing and misfire limits must be considered to operate in HCCI mode 1 . 

• Ringing limit = 5 MW/m 2 

• COV of IMEP (misfire) limit = RD387 5% Gasoline 

470 

450 

430 

410 

390 

370 

350 

330 

310 

290 

270 

7.0 bar EMEP 

7.4 bar EMEP 

7.8 bar EMEP 

8.2 bar EMEP 

8.6 bar EMEP 

9.0 bar EMEP 

9.4 bar EMEP 

9.8 bar EMEP 

fuel energy ( J / cycle ) 

EMEP ( bar) 

 

displaced volume( 

L)*100 

-2 0 2 4 6 8 10 12 14 


1 

Hagen, L., Bohac, S., Lavoie, G., Assanis, D., “Fuel Effects on Heat Release and Maximum Load with HCCI in the UM FFVA Engine,” presented 

at Sandia National Lab, Livermore CA, February 2011. 


Combining HCCI Ringing and Misfire Limits 

IMEP g [kPa] 

470 

450 

430 

410 

390 

370 

350 

330 

310 

290 

270 

7.0 bar EMEP 

7.4 bar EMEP 

7.8 bar EMEP 

8.2 bar EMEP 

8.6 bar EMEP 

9.0 bar EMEP 

9.4 bar EMEP 

9.8 bar EMEP 

RD387 Gasoline 

Ringing 

Limit 

-2 0 2 4 6 8 10 12 14 


Stability 

Limit 

* 

Hagen, L., Bohac, S., Lavoie, G., Assanis, D., “Fuel Effects on Heat Release and Maximum Load with HCCI in the UM FFVA Engine,” presented 

at Sandia National Lab, Livermore CA, February 2011. 


Emissions from HCCI 



PM 

HC 

NO x 

CO 2 

CO 



We’ll consider three groups of pollutants: 

• NO x emissions 

• Mono-nitrogen oxides (NO + NO 2 ) 

• NO x contributes to smog, acid rain, and respiratory ailments 

• THC emissions 

• Total hydrocarbons 

• Reactive hydrocarbons contribute to the formation of ground-level 

ozone, which is the main component of photochemical smog. 

• Some HC are carcinogenic (e.g., benzene, 1,3-butadiene) 

• Methane contributes to climate change 

• PM emissions 

• Particulate matter 

• Solid or liquid particles suspended in a gas 

• Main components are elemental carbon (EC) and organic carbon (OC) 

• PM causes inhalation health effects (lung irritant and desorption of OC) 



NO x emissions 

• Measured using Fourier Transform 

InfraRed detector (FTIR) 

• MKS 2030HS 

FTIR 

FID 

THC emissions 

• Measured using Flame Ionization 

Detector (FID) 

• Horiba FIA 236 



Smoke Meter 

PM emissions 

• EC measured using Smoke Meter and EC correlation 1 

• AVL 415S 

• Total PM measured with Partial Flow Dilution Tunnel 

and Gravimetric Filter Analysis 

• In-house designed dilution tunnel, Pall Emfab 

TX40 filters, Sartorius microbalance 

Mini Dilution Tunnel (MDT) 

Loaded PM Filter 

Filter Conditioning Glove Box 

1 

Christian, V. R., Knopf, F., Jaschek, A., Schneider, W., 1993, “Eine Neue Messmethodik der 

Bosch-Zahl mit Erhoehter Emphfindlichkeit”, MTZ, 54(1), 16-22. 



2.0L HCCI Engine 

Multi-Cylinder HCCI Test Engine 

MY2010 GM Ecotec LNF 2.0L DISI turbocharged I4 

Side-mounted DI injector 

HCCI modifications: 

• CR increased from 9.25 to 11.25 

• Low-lift HCCI camshafts with NVO 

• Small Borg Warner Turbocharger 

• Small Eaton supercharger 

Multi-Cylinder Diesel Test Engine 

MY2008 Ford Powerstroke 6.4L V8 diesel 

Common rail fuel injection 

6.4L Diesel Engine 



NO x Emissions from HCCI 

With lean HCCI combustion where < 0.6 (Lambda > 1.6), NO x 

emissions can be maintained below 10 ppm and NO x 

aftertreatment is not needed. 

The challenge comes with moderately lean SACI or SI (0.60 < < 

1.00) 



HC Emissions from HCCI 

Hydrocarbons can be relatively high, especially when high 

amounts of residual gas are used to initiate combustion at low 

load. 

At 2000 rpm, 3 bar BMEP, HCCI increases THC from 1500 to 2500 

ppmC 1 . 


PM (mg/m3) 

PM Emissions from HCCI 

40 

35 

30 

25 

20 

15 

10 

5 

0 


OC Mass 

EC Mass 

>40 >40 

>40 

diesel 6 bar diesel 9 bar HCCI 60°BTDC HCCI 52°BTDC HCCI 50°BTDC 

Diesel 

1500 rpm, 6 bar / 2500 rpm, 9 bar 

SOI=3.5ºATDC 

14% EGR 

HCCI 

Wall-guided spray 

Split injection, 50/50 mass ratio: 

SOI of 1 st injection = 345ºBTDC 

EOI of 2 nd injection = 60º, 52º, 50ºBTDC 

1500 rpm, 1.9 bar BMEP 

• Injection at 50°BTDC produces diesel-like EC…particulate filter needed. 

• All injection timings displayed excessive OC in PM…TWC may oxidize this. 


Summary 


Summary 

• Significant efficiency benefits can be attained by increasing gamma 

(γ=C p /C v ). 

• LTC increases gamma by reducing temperature and operating lean. 

• Boosting moves efficiency peak to leaner mixtures. 

• Different requires different combustion modes: 

‣SI: = 0.65 - 1.0 

‣HCCI: = 0.20 – 0.45 

‣SACI: = 0.45 – 0.65 

• As HCCI load increases, ringing and misfire limits converge. 

• Lean HCCI has low NO x , high THC, and can have high PM 

• In moving from stoichiometric SI to lean HCCI, we improve 

efficiency but make combustion control and emissions control 

more challenging. 


Future challenges 



1. Optimization of direct injection for HCCI 

– Multiple injections to avoid spray-wall impingement, pool files and PM 

– Charge stratification to reduce ringing and increase load limit 

– Injection during NVO to generate HC radicals and promote combustion 

2. Developing effective after-treatment for HCCI 

– Low temperature, high THC, different HC composition 

– Improved THC storage of NO x and HC for mode switches 

– PM control with TWC or DPF 

3. Control of HCCI combustion and mode switching 

– Combustion feedback (Pcyl, ion probe, crankshaft acceleration, exhaust 

lambda and NO x sensors) 

– Transients and mode switches 

4. Engine design for HCCI 

– High cylinder pressure, knocking/ringing, audible noise 


Part 2: Diesel 



Outline 

• Why consider low temperature combustion (LTC) for diesel engines 

• Reducing emissions with diesel LTC 

• Experimental setup for studying diesel LTC 

• Implementation of LTC on a diesel engine 

• A closer look at diesel LTC hydrocarbon emissions 

• Summary 

• Future challenges 



for diesel engines 


Conventional Diesel Combustion 

Good efficiency as compared to gasoline engines (high γ, high CR, no 

throttle, boosted) 

2009 VW 2.0L TDI Diesel: 265 g/kWh @ 2000 rpm, 3 bar BMEP 18% ↓ 

2010 GM 2.0L LNF Gasoline: 323 g/kWh @ 2000 rpm, 3 bar BMEP 

Good control of combustion 

Combustion phasing is set with injection timing 

Poor emissions characteristics 

High NO x and PM (heterogeneous mixture) requires expensive and complex aftertreatment 

US EPA Tier 2, Bin5 

Euro 5 

DOC + cDPF + SCR with used by VW, Mercedes, GM 

DOC + cDPF on small cars 

DOC + cDPF + SCR or LNT on heavy cars 



for diesel engines 

To improve emissions while maintaining efficiency. 


Reducing emissions with diesel LTC 


Phi – T Diagram 

Conventional diesel combustion straddles the soot and NO formation zones. 



Soot 

1600-2400K 

Ideal temperature range for pyrolysis of fuel into soot nucleation sites 

2400K 

Fuel decomposes into gaseous HC fragments rather than pyrolyzing into soot 

NO x 

>2100K 

NO formation occurs 

highly temperature dependent 

1 

Kamimoto, T., Bae, M., “High Combustion Temperature for the Reduction of Particulate in Diesel Engines,” SAE 880423, 1988. 



By increasing mixing (to reduce local phi) and reducing temperature, LTC can greatly 

reduce both soot and NO. 

LTC can be achieved by using heavy 

EGR (40-50%), earlier injection timing, 

and lower CR. 

NO x ↓ by lower temperatures. 

PM ↓ by lower temperature and more 

premixed combustion (lower local phi). 

But CO and HC ↑ and must be 

addressed. 


Diesel LTC Concept 

Reduce PM and NO x emissions with LTC 1,2 . 

Use DOC to oxidize the increased CO and HC 2 . 

LTC 

DOC (oxidize CO and HC) 

+ 

1 

Jacobs, T. J., Bohac, S. V., Assanis, D. N., Szymkowicz, P. G., 2005, “Lean and Rich Premixed Compression Ignition Combustion in a Light-Duty 

Diesel Engine,” SAE Paper 2005-01-0166, SAE Transactions – Journal of Engines, 114. 

2 

Bohac, S. V., Han, M., Jacobs, T. J., López, A. J., Assanis, D. N., Szymkowicz, P., 2006, “Speciated Hydrocarbon Emissions from an Automotive 

Diesel Engine and DOC Utilizing Conventional and PCI Combustion,” SAE Paper 2006-01-0201, SAE Transactions – Journal of Fuels and 

Lubricants, 115. 


Experimental setup for studying diesel LTC 


Experimental Setup 

Engine 

Manufactured by General Motors / Isuzu 

1.7L 4-cylinder automotive diesel engine 

B/S = 79/86mm 

4 valves per cylinder with swirl control valves 

Common rail fuel injection 

Turbocharger with VGT, intercooler 

Intake throttle 

EGR cooler 

16:1 compression ratio (lowered from 19:1) 

Multi-Cylinder LTC Research Engine 

Platinum-based production DOC 

0.50 L 

400 cells/in 2 

Cordierite ceramic monolith 

Alumina and silica washcoat 

Pt:Pd:Rh = 1:0:0 

150 g/ft 3 platinum 

Platinum-Based DOC 


Emissions instrumentation 

Emissions bench (AVL CEB II) 

NO x , CO, CO 2 , O 2 , THC 

Experimental Setup 

Emissions Bench 

Smoke Meter 

Filter smoke meter (AVL 415S) 

FSN and calculated soot 1 

Gas Chromatography for volatile HC (C 1 -C 8 ) 

Dilution and Tedlar bag used for sampling 

Shimadzu GC-17A with FID for analysis 

HC Speciation Laboratory 

Gas Chromatography for semi-volatile HC (C 6 -C 20 ) 

Tenax trap and MFC used for sampling 

Tekmar Aerotrap 6000 used for desorption 

Shimadzu GC-17A with FID for analysis 

1 

Christian, V. R., Knopf, F., Jaschek, A., Schneider, W., 1993, “Eine Neue 

Messmethodik der Bosch-Zahl mit Erhoehter Emphfindlichkeit”, MTZ, 54(1), 16-22. 


Implementation of LTC on a diesel engine 


Implementation of LTC on a Diesel Engine 

Operating condition and development targets: 

Lean 

Conventional 

Lean 

LTC 

Speed (RPM) 1500 1500 

BMEP (bar) 3.5 to 4 3.5 to 4 

Rail Pressure (bar) 300 1000 

EGR Rate (%) 32 41-45 

Inj Timing (°BTDC) 3-15 9-18 

EI-NO x (g/kg-fuel)


32% EGR 32% EGR 

• NO x decreases as injection timing is retarded and combustion temperature is 

reduced. 

• PM first increases then decreases with injection timing. 

– Initial increase in PM is due to shorter ignition delay and increased diffusion burn. 

– Subsequent decrease in PM results from more mixing and lower temperature. 



32% EGR 32% EGR 

• NO x decreases as injection timing is retarded and combustion temperature is 

reduced. 

• PM first increases then decreases with injection timing. 

– Initial increase in PM is due to shorter ignition delay and increased diffusion burn. 

– Subsequent decrease in PM results from more mixing and lower temperature. 



32% EGR 32% EGR 

Advanced injection: 

• Increases NO x 

• Increases noise 

• Decreases BSFC 


Lean Conventional Combustion that Meets Targets 

Lean 

Conventional 

Speed (RPM) 1500 

BMEP (bar) 3.96 

Rail Pressure (bar) 300 

EGR Rate (%) 32 

Inj Timing (°BTDC) 5 

EI-NO x (g/kg-fuel) 4.28 

EI-PM (g/kg-fuel) 0.34 

Noise (dB) 88.2 

BSFC (g/kW-hr) 235 



• Retarded injection reduces NO x . 

• Increased EGR greatly reduces NO x . 



EI-PM 

LTC 

• Higher EGR increases PM at advanced injection timing. 

• Higher EGR with retarded injection timing reduces PM. Here both NO x and PM 

decrease with retarded injection timing. 



• Increasing EGR from 42% to 45% significantly retards combustion. 

• Retarding injection from 15 to 9°BTDC also significantly retards combustion. 



• Noise target < 90 dB 

• BSFC target < 5% efficiency loss relative to conventional combustion. 



Lean 

Conventional 

Lean PCI 

Speed (RPM) 1500 1500 

BMEP (bar) 3.96 3.75 

Rail Pressure (bar) 300 1000 

EGR Rate (%) 32 45 

Inj Timing (°BTDC) 5 15 

EI-NO x (g/kg-fuel) 4.28 0.31 

EI-PM (g/kg-fuel) 0.34 0.07 

Noise (dB) 88.2 88.7 

BSFC (g/kW-hr) 235 246 

Jacobs, T. J., Bohac, S. V., Assanis, D. N., Szymkowicz, P. G., 2005, “Lean and Rich Premixed Compression Ignition Combustion in a Light-Duty 

Diesel Engine,” SAE Paper 2005-01-0166, SAE Transactions – Journal of Engines, 114. 


A closer look at diesel LTC Hydrocarbon emissions 


Operating Conditions 

Lean Conv. Lean LTC Rich LTC 

Speed (rpm) 1500 1500 1500 

BMEP (bar) 3.9 3.9 3.6 

Inj.Timing(°BTDC) 5° 15° 25° 

Rail P (bar) 300 1000 1000 

Comb. A/F (-) 28.7 16.2 12.5 

EGR (%) 33 45 49 

Exhaust T (°C) 290 283 254 

Comb. Noise (dB) 88 89 89 

BSFC (g/kWh) 242 249 313 

• low PM and NO x 

• DOC oxidizes CO and HC 

• low PM and NO x 

• generate CO and H 2 

for LNT regeneration 


Engine-Out Emissions 

* 





Engine-Out HC Composition by Carbon Number 


Engine-Out HC Composition by Carbon Number 


Late-Forming Species (Engine-Out) 

Rich PCI 

Lean Conv. 

Lean PCI 


DOC Conversion Efficiency 


DOC Conversion Efficiency 


Why Does the DOC Perform Poorly with Rich LTC 

• Rich LTC has low O 2 concentration 

• There is insufficient O 2 to oxidize all of the CO and HC. 

• But the DOC does not even use the O 2 that is available 

(pre-DOC and post-DOC O 2 concentrations are both 0.9%). 

• Injecting O 2 ([O 2 ]=2%) does not improve CO and HC conversion 

efficiencies. 



• Rich PCI has low O 2 concentration – No. 

• Rich LTC has low exhaust temperature 

• Rich LTC exhaust is colder (254ºC) than lean LTC exhaust. 

• But the DOC is active if lean LTC exhaust temperature is reduced to 

220ºC. 

• And the DOC is inactive if rich LTC exhaust temperature is increased to 

320ºC. 




• Rich PCI has low exhaust temperature – No. 

• Rich LTC has different hydrocarbon properties (i.e., less reactive HC) 

• Rich LTC has a higher fraction of C 1 -C 2 hydrocarbons. 

• But even conversion of reactive species like CO, acetylene and olefins 

is severely reduced. 





• Rich PCI has different hydrocarbon properties (i.e., less reactive HC) – No. 

• Rich LTC has a high CO/O 2 ratio 

• Platinum preferentially adsorbs CO over O 2 at temp. typical of diesel 

exhaust. 

• If p CO ≥ p O2 then the DOC’s active sites become saturated with CO, 

leaving no room for O 2 chemisorption. This would cause CO and HC 

oxidation to stop. 

• Lean LTC CO/O 2 = 0.29 

• Rich LTC CO/O 2 = 6.0 





• Rich PCI has different hydrocarbon properties (i.e., less reactive HC) – No. 

• Rich PCI has a high CO/O 2 ratio – Yes. 

• To make DOC operational with rich LTC: 

• Higher exhaust temperature to increase CO/O 2 threshold, as in SI 

engines. 

• Different precious metal or DOC formulation. 

* 





Summary 


Summary 

• Diesel LTC simultaneously reduces PM and NO x emissions to very 

low levels. It does this by increasing mixing and reducing 

combustion temperature. 

• Greater mixing and lower temperature increases HC and CO 

emissions, necessitating the use of a DOC. 

• Rich LTC used with a DOC can be used to generate CO and H 2 to 

regenerate an LNT, but high CO/O 2 exhaust ratios can easily 

deactivate a platinum DOC. 




Future Challenges 

• Combustion feedback sensors and controls. This is especially 

important in the marketplace where different fuel properties and 

ambient temperatures are common. 

• DOC that are active at low temperatures and with high levels of CO 

and HC. 


Thank you for your attention! 

E-mail: sbohac@umich.edu 


Acknowledgements 

• Professor Jan Macek and the Josef Bozek Research Centre at 

the Czech Technical University in Prague 

• Dr. George Lavoie, Dr. Jason Martz, Dr. Aris Babajimopoulos, 

Elliot Ortiz-Soto, Prof. Dennis Assanis – Gasoline LTC 

simulations 

• Jiri Vavra, Luke Hagen, Weiyang Lin, Jianye Su – Gasoline LTC 

experiments 

• Prof. Tim Jacobs, Prof. Manbae Han, Alberto Lopez – Diesel LTC 

experiments 


Acknowledgements - Projects 

• Marie Curie VECOM Project 

• A University Consortium on Efficient and Clean High Pressure 

Lean Burn (HPLB) Engines; U.S. Department of Energy 

• Advanced Controllable Combustion Enabling Systems and 

Solutions (ACCESS) for High Efficiency Light Duty Vehicles; 

Bosch and U.S. Department of Energy 

• General Motors – University of Michigan Collaborative 

Research Laboratory (GM-CRL); General Motors 


Thank you for your attention! 

E-mail: sbohac@umich.edu 


Selected Diesel LTC References 

Northrop, W. F., Assanis, D., Bohac, S., 2011, “Evaluation of Diesel Oxidation Catalyst Conversion of Hydrocarbons and Particulate Matter from Premixed Low Temperature Combustion of Biodiesel,” SAE Paper 

2011-01-1186, SAE Int. J. Engines 4(1):1431-1444. 

Northrop, W. F., Bohac, S. V., Chin, Jo-Yu, Assanis, D. N., 2011, “Comparison of Filter Smoke Number and Elemental Carbon Mass from Partially Premixed Low Temperature Combustion in a Direct Injection Diesel 

Engine,” Journal of Engineering for Gas Turbines and Power, 133(10), pp. 102804-1 to 102804-6. 

Han, D., Ickes, A., Bohac, S. V., Zhen, H., Assanis, D. N., 2011, “Premixed Low-Temperature Combustion of Blends of Diesel and Gasoline in a High Speed Compression Ignition Engine,” Proceedings of the 

Combustion Institute, 33: 2, 3039-3046. 

Northrop, W. F., Madathil, P. V., Bohac, S. V., Assanis, D. N., 2011, “Condensational Growth of Particulate Matter from Partially Premixed Low Temperature Combustion of Biodiesel in a Compression Ignition 

Engine,” Aerosol Science and Technology, 45: 1, 26-36. 

Han, D., Ickes, A., Assanis, D. N., Zhen, H., Bohac, S. V., 2010, “The Attainment and Load Extension of High-Efficiency Premixed Low-Temperature Combustion with Dieseline in a Compression Ignition Engine,” 

Energy & Fuels, 24, pp. 3517-3525. 

Northrop, W. F., Vanderpool, L., Madathil, P., Assanis, D. N., Bohac, S. V., 2010, “Investigation of Hydrogen Emissions in Partially Premixed Diesel Combustion,” Journal of Engineering for Gas Turbines and Power, 

132(11), pp. 112803-1 to 112803-6. 

Ickes, A. M., Bohac, S. V., Assanis, D. N., 2009, “Effect of 2-Ethylhexyl Nitrate Cetane-Improver on NO x Emissions from Premixed Low-Temperature Diesel Combustion,” Energy & Fuels, 23, pp. 4943-4948. 

Ickes, A. M., Bohac, S. V., Assanis, D. N., 2009, “Effect of Fuel Cetane Number on a Premixed Diesel Combustion Mode,” Int. J. Engine Res., 10(4), pp. 251-263. 

Northrop, W. F., Bohac, S. V., Assanis, D. N., 2009, “Premixed Low Temperature Combustion of Biodiesel and Blends in a High Speed Compression Ignition Engine,” SAE Paper 2009-01-0133, SAE Int. J. Fuels Lubr. 

2(1):28-40. 

Han, M., Assanis, D. N., Bohac, S. V., 2009, “Sources of Hydrocarbon Emissions from Low-Temperature Premixed Compression Ignition Combustion from a Common Rail Direct Injection Diesel Engine,” Combustion 

Science and Technology, 181, pp. 496-517. 

Han, M., Assanis, D. N., Bohac, S. V., 2008, “Characterization of Heat-up Diesel Oxidation Catalysts through Gas Flow Reactor and In-situ Engine Testing,” Proc. IMechE Part D – Journal of Automobile Engineering, 

222(9), pp. 1705-1716. 

Han, M., Assanis, D. N., Bohac, S. V., 2008, “Comparison of HC Species from Diesel Combustion Modes and Characterization of a Heat-up DOC Formulation,” International Journal of Automotive Technology, 9(4), 

pp. 405-413. 

Han, M., Assanis, D. N., Jacobs, T. J., Bohac, S. V., 2008, “Method and Detailed Analysis of Individual Hydrocarbon Species from Diesel Combustion Modes and Diesel Oxidation Catalyst,” Journal of Engineering for 

Gas Turbines and Power, 130(4), pp. 042803-1 to 042803-10. 

Northrop, W., Jacobs, T., Assanis, D., Bohac, S., 2007, “Deactivation of a Diesel Oxidation Catalyst due to Exhaust Species from Rich Premixed Compression Ignition Combustion in a Light-Duty Diesel Engine,” Int. J. 

Engine Res., 8(6), pp. 487-498. 

Knafl, A., Busch, S. B., Han, M., Bohac, S. V., Assanis, D. N., Szymkowicz, P. G., Blint, R. D., 2006, “Characterizing Light-Off Behavior and Species-Resolved Conversion Efficiencies during In-Situ Diesel Oxidation 

Catalyst Degreening,” SAE Paper 2006-01-0209, SAE Transactions – Journal of Fuels and Lubricants, 115.

VECOM Low Temperature Combustion

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