04S-11

Combustion in a Centimeter-Scale Four-Stroke EngineJoseph Papac and Derek Dunn-Rankin

University of California Irvine

Abstract

Centimeter-scale liquid hydrocarbon fueled engines show promise for future use as high powerdensity mobile power sources. Experimental results have shown that the peak power density of suchengines is approximately 300 W/kg. For comparison, most electrochemical devices are limited topower densities of approximately 100 W/kg. While the power density of these devices is impressive,their operation is marked by a high percentage of unburned hydrocarbons and thus a low fuel con-version efficiency. The fuel-laden exhaust gives the impression of rich operation, however the systemoperates under overall near-stoichiometric conditions. We believe that two modes of combustiontake place. An unburned wall film develops in the cylinder and acts as a cooling mechanism. Fuelevaporation from this film produces a relatively rich mixture near the walls that might sustain adiesel-like combustion process. Away from the walls, the mixture burns in a homogeneous chargecompression ignition (HCCI) mode. Cylinder pressure measurements show that the compressionprocess is far from ideal, mainly resulting from leakage past the piston ring.

Introduction

Defining the power requirements for autonomous systems has two principal components —energy density (i.e., energy per unit mass, driven by the desired operating duration without refu-eling) and power density (driven by the maximum power needed for the application). Ideally bothhigh energy and high power density are desirable, but often there is a trade-off between these two.Figure 1 shows this trade-off schematically by plotting power density versus energy density (on amass basis) for a variety of typical power sources. The power sources shown represent the behaviorof operating systems, i.e. not projected performance, but actual performance of thermochemicaland electrochemical power devices. The diagonal lines on the plot define operating duration. Theengine curves have two asymptotes controlled by the ratio of fuel mass to system mass. When itis carrying very little fuel, the mass of the system is dominated by the power conversion structure,which leads to very low energy density for the system. At the opposite extreme, virtually all ofthe system mass is fuel and so the energy density is constrained primarily by the system’s thermalefficiency, appearing as a vertical asymptote. Also noted on the curve is the energy density ofa typical liquid hydrocarbon fuel at approximately 12,000 Watt-hr/kg. For an engine with 25%thermal efficiency, the asymptote is near 3,000 Watt-hr/kg. In this fairly approximate analysis, therelationship between fuel mass and the mass of the structure to contain the fuel is neglected. Inthe case of a compressed gaseous fuel this assumption would not be valid, but our focus is on liquidhydrocarbon fuels because of their superior volumetric energy density.

The figure also shows the challenge of using electrochemical devices for applications that demandhigh power density. For any application where more than 100 W/kg is needed (all aircraft andmost vehicles, for example) the only choices with decent operating duration are liquid hydrocarbonfueled engines. There is some expectation that fuel cells and advanced battery designs (nickel-metal

1

10000

Sour

ce p

ower

den

sity,

W/k

g

1 101

100 100001000

10

100

1000

100 hr6 min

10 hr

1 hr

36 sec

3.6 sec

lead acid batterymini-diesel

automobileengines

Hydrocarbon fuel > 10000 Whr/kg

primarylithium

rechargeablelithium

modelairplaneengine

fuel cell

high powerlead acid

Stored energy density, Whr/kg

Figure 1: Ragone plot of source power density vs. stored energy density of power devices: fuel cell(solid curve), batteries (dotted curves), and combustion engines (dashed curves).

hydride and lithium-based) will eventually achieve sufficient power densities to be considered, butthere are many challenges. For example, since the energy release in these electrochemical systemsoccurs at a surface, power performance in these devices is governed by surface area. In addition, therapid transport of electrons and ions from deep within a storage matrix produces resistive losses.Hence, as the power demands increase, energy density capabilities decrease. Volumetric energyrelease through a process like combustion will be needed for maximum power density.

It was the recognition of the energy considerations discussed above that generated an enthu-siastic search for small scale engines to provide high power density in tiny packages e.g. [1, 2].After considerable study and effort [3–8], it appears that the smallest practical combustion enginesare likely to settle in the range of a centimeter or so in their critical dimension. A combinationof surface-to-volume challenges (primarily thermal management, friction, and combustion reactiontime) and the realization that electrochemical devices are often reasonable alternatives for tinypower outputs has pushed attention to engines in the range of 30–1,000 Watts. As an example,approximately one kg of fuel (or approximately one liter of fuel) would be required for a 10%efficient 100 W shaft output engine to operate for 10 hours. If we are already carrying around1,000 cc of liquid fuel, whether the engine is 1 cc or 0.1 cc is not likely to be significant. Plentyof challenges remain, even for miniature engines of centimeter size, and this paper examines someof those challenges through the experimental evaluation of a commercially available small scaleengine. In particular, we examine an engine designed for radio controlled airplanes. Engines forradio controlled vehicles (airplanes and automobiles) have had many years of development timeand they have hence evolved considerably [9]. This evolution helps assure a performance that has

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optimized the engine to meet the demands (cost, power, and efficiency) of the application for whichthey are designed.

Description of the Engine

The O.S. Engines FS-30-S has been chosen for this study because it is a representative exampleof a modern mass-produced model engine design and because it mimics a full-scale automobileengine before emission reduction demanded the major technological developments of fuel injectionand automatic engine control. The engine is a single cylinder, 4.89 cc displacement four-strokedesign, with single intake and exhaust valves driven by pushrods. The piston has a single pistonring. Fuel is delivered to the engine by a needle valve carburetor. The fuel tank is pressurizedby the high temperature exhaust gas. The pressure driving the fuel through the carburetor isinfluenced by the hydrostatic pressure of the liquid fuel and by the temperature and pressure ofthe exhausting products. While this design has the advantage of simplifying the balance of plant,it contaminates the fuel supply and can produce inconsistency in the fuel-air mixture delivered tothe engine. Ignition is initiated by a resistively heated glow plug wire with a platinum catalyst.Once the engine has reached a steady state operating temperature, electrical energy to the glowplug is no longer necessary. The retained heat of the glow plug continues to provide a catalytic hotspot for ignition within the engine cylinder.

Lubrication of the engine parts is done by premixing the fuel with oil. This simplifies the enginedesign, but all of the unburned oil is exhausted into the atmosphere and its high vapor pressureresults in deposition on surfaces in the immediate surroundings of the engine.

Full scale reciprocating engines have power density performance very close to that both desirableand necessary for autonomous applications. With a few simple assumptions, it is possible toidentify a volumetric power expectation for these engines operating on methanol as does the modelengine. With combustion reaction timescales on the order of one millisecond, thermal efficiencies ofapproximately 20%, and methanol fuel energy density on the order of 20 MJ/kg, we can expect atmost 650 Watts/cc of engine displacement. This value assumes that there is no mixing time requiredand that there are no pressure drops between the intake and engine cylinder. The manufacturerof the FS-30-S engine reports approximately 1/4th this value, suggesting that, for reasons alreadymentioned, the efficiency is closer to 5% for small scale engines.

Thermodynamic Analysis

Experimental studies have shown that the compression and expansion processes in spark-ignited internal combustion engines are well fitted by a polytropic relation [10]. The value of thepolytropic exponent for typical fuels is n = 1.3 ± 0.05. While the method of ignition differs ina glow-ignited IC engine, the compression and expansion processes are the same. Therefore, weassume that the polytropic relation holds.

For simplicity, we assume that the incoming fuel/air mixture is a fully premixed ideal gascomposed of methanol vapor (CH3OH) and air (21% by volume O2 and 79% N2) in stoichiometricproportions at 350 K (slightly preheated) and 1 atm. The temperature, T2, and pressure, P2, ofthe mixture at top-dead-center (TDC) prior to combustion are calculated by equations (1) and (2).

T2

T1=

(Vc

Vd + Vc

)n−1

(1)

3

P2

P1=

(Vc

Vd + Vc

)n

(2)

where Vc = 0.81 cc is the clearance volume and Vd = 4.89 cc is the displaced volume. With theassumption of n = 1.3, the temperature and pressure are calculated: T2 = 629 K and P2 = 12.7atm.

Turbulence in the Centimeter Scale

In full size spark ignited engines the initial phase of combustion involves a relatively smoothspherical laminar flame. As the flame grows, the flame front becomes increasingly distorted by theturbulent flow field through which it is propagating and develops a highly wrinkled and multiplyconnected structure [11]. Abraham, Williams, and Bracco [12] suggest that turbulence intensitiesin spark-ignition engines can be approximated by v′rms ≈ vp/2 (at the time of spark), where vp

is the piston velocity. Also, the integral scale can be approximated by l0 ≈ h/2, where h isthe instantaneous clearance between the top of the piston and the cylinder head in disk-shapedcombustion chambers. From geometrical and kinematic analyses, the instantaneous piston speedin m/s is related to the rotational speed, N (rev/s), the crank angle after TDC, θ, the ratio of theconnecting rod length to crank radius, R∗, and the engine stroke, L, by equation (3).

vp = 2LNπ

2sin θ

[1 +

cos θ

(R∗2 − sin2 θ)1/2

](3)

Also, the instantaneous clearance height, h, is given by equation (4),

h

hTDC= 1 +

12(rc − 1)[R∗ + 1− cos θ − (R∗2 − sin2 θ)1/2] (4)

where rc is the compression ratio. The dimensions of the O.S. engine were measured: hTDC = 2.67mm, rc = 7.05, L = 16.4 mm, and R∗ = 3.49. The piston velocity is zero at the beginning of thestroke, reaches a maximum near the middle of the stroke, and decreases to zero at the end of thestroke. Approximate values of the turbulence intensity, v′rms = 2.68 m/s, and the integral scale,l0 = 1.73 mm, are calculated with N = 166.7 rev/s (10,000 rpm) and θ = 30 before TDC—atypical value for spark timing in automotive engines.

The turbulence Reynolds number based upon the integral scale, Rel0 , is defined by equation(5).

Rel0 =ρv′rmsl0

µ(5)

The dynamic viscosity, µ = 3.15 × 10−5 N·s/m2, and the density, ρ = 7.11 kg/m3, of the workingfluid are assumed to be equal to that of air at T2 = 629 K and P2 = 12.7 atm. Under theseapproximations Rel0 = 1048 for this engine.

The laminar burning velocity, SL = 119 cm/s, of the stoichiometric methanol/air mixture at629 K and 12.7 atm is calculated from the empirical correlation developed by Metghalchi and Keck[13], shown in equation (6),

SL = SL,0

(T2

T0

)α(P2

P0

)β

(6)

where SL,0, α, and β are dependent on the fuel type and equivalence ratio, and T0 = 298 K andP0 = 1 atm are the reference temperature and pressure.

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The flame thickness, δL, is calculated from Spalding’s laminar 1-D premixed flame approach[14],

δL =2α

SL(7)

where α = 0.241 cm2/s is the thermal diffusivity of air. The value of α used is that of air at themean temperature of T = (T + Tf )/2, where T = 629 K, Tf = 2, 000 K and pressure of 12.7 atm.The flame thickness is 40.4 µm. The Dahmkohler number (Da), which relates the chemical timeto the flow time, is defined by equation (8).

Da =(

l0δL

)(SL

v′rms

)(8)

The values of Da = 19.1 and Rel0 = 1048 are consistent with those typical of full size IC engines.They lie near the boundary between the wrinkled laminar flame and the flamelets-in-eddies regimesof turbulent combustion [15]. The effects of turbulent flow in the centimeter scale appear to be thesame as in full size internal combustion engines.

Residence Time

With the previously stated assumptions, we can estimate the turbulent burning velocity. Forthis calculation, we will use the wrinkled laminar flame correlation developed by Klimov [16] andshown in equation (9).

ST

SL= 3.15

(v′rms

SL

)0.7

(9)

The obtained turbulent burning velocity is ST = 7.37 m/s. The time required for flame prop-agation across the cylinder may be estimated. We assume that the flame kernel is ignited at thecenter of the combustion chamber and the flame propagates radially outward to the walls at theconstant rate of the turbulent burning velocity. Then the time required for complete combustionis 1.3 ms.

If we assume that the flame propagation begins at 30 before TDC and the rotational speedis 10,000 rpm (near the upper limit of engine operation), then the total crank angle over whichcombustion occurs is approximately 80. Equation (4) may by solved with θ = 50 to obtain thepiston position at which complete combustion occurs. The piston position is 3 mm below TDC.This results in a relatively minor increase in cylinder volume compared to the change in volumeover the entire stroke (0.9 cc as compared to 4.9 cc).

This analysis shows that there is sufficient residence time for complete combustion occur.

Experimental Setup

An electric motor dynamometer is used for power measurement. The rotational speed of theengine shaft, ω, is measured optically by determining the frequency of the reflections of a helium-neon laser. The light is reflected off of the engine shaft once per crankshaft revolution and focusedonto a fast photodiode, which outputs a periodic waveform corresponding to the angular velocityof the engine. A measured load is applied to the engine via a belt and pulley system connectedto an Ever Motor ERS-380PM-3270 permanent magnet (PM) DC electric motor. Transmissionlosses in the belt are assumed and qualitatively confirmed to be negligible. Due to the interaction

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between the stator field and the armature current, the torque generated by a PM DC electricmotor is directly proportional to the current [17]. The value of the torque constant, kT , is 5.16milli-Newton-meters/amp. The electric motor is connected in parallel to a high power resistor bank.The external resistance, Rext, can be varied between 0.6 Ω and 10.0 Ω. An increase in externalresistance reduces the applied load by decreasing the current provided to the motor windings. Thevoltage, V , between the electric motor’s positive and negative terminals is measured. From thisvoltage measurement the torque, τ , of the engine can be calculated from equation (10),

τ =D2 · kT · VD1 ·Rext

(10)

where D2 and D1 are the diameters of the pulleys attached to the engine and the electric motorrespectively. The mechanical power of the engine is the product of angular velocity and torque,τ = ω · P .

The fuel flow rate is determined by taking a series of measurements of the mass of the fuel inthe fuel tank once the engine is operating at a steady state condition. The fuel mass flow rate isthen calculated from a least squares fit of the data. The intake air volumetric flow rate is measuredby a bubble-meter calibrated rotameter. To prevent thermal failure of the engine components, a 77CFM fan provides external cooling. Steady state operation is achieved over a two minute warm-upperiod before data is acquired. A fixed load setting is applied to the engine and the rotationalspeed is held constant while measurements are logged over a 10 minute interval.

Cylinder pressure measurements are obtained with an Omega model DPX101-500 pressuretransducer mounted to the cylinder head.

Results

The engine was tested with three fuel combinations. Fuel mixture A is composed of 79%methanol, 3% nitromethane (CH3NO2), and 18% castor oil (the manufacturer’s specified minimumvalue). Fuel mixture B is composed of 72% methanol, 10% nitromethane, and 18% castor oil. Fuelmixture C is composed of 62% methanol, 20% nitromethane, and 18% castor oil. The lower heatingvalue (QLHV ) of methanol is 20.0 MJ/kg and the QLHV of nitromethane is 11.3 MJ/kg. Thus, itis apparent that the energy density of a fuel mixture decreases as the nitromethane concentrationincreases (relative to methanol concentration). The advantage of adding nitromethane to the fuelmixture is that the ratio of chemical energy to volume of the stoichiometric reactant mixtureincreases. The nitromethane molecule contains two oxygen atoms while the methanol moleculecontains only one oxygen atom. Nitromethane, therefore, acts as a superior oxygen carrier, requiringless air for combustion than does methanol, thereby allowing a greater volume of fuel into the samesize combustion chamber. The stoichiometric molar air-fuel ratios (A/F) of pure methanol and purenitromethane are 7.14 and 3.57 respectively and the volumetric energy density of a nitromethane-air mixture is nearly double that of a methanol-air mixture (6.73 versus 3.51 J/cm3 gas mixture atstandard temperature and pressure). Because fuel evaporation is important for engine cooling, thehigh latent heat of methanol is another reason for its choice as a model engine fuel.

The engine was operated over a range of loads (8–60 milli-Newton-meters) and engine speeds(3,500–13,500 rpm). Although we measured many engine properties during the tests, we concen-trated on the engine’s power and efficiency performance. The fuel conversion efficiency, ηf , is given

6

0

10

20

30

40

50

60

70

80

90

0% 2% 4% 6% 8% 10% 12%Efficiency

Brak

e po

wer

(wat

ts)

Mix AMix BMix C

Figure 2: Plot of brake power vs. fuel conversion efficiency

by equation (11),

ηf =Wb

mf (YmethanolQLHV,methanol + YnitromethaneQLHV,nitromethane(11)

where Wb is the measured brake power, mf is the fuel mass flow rate, Yi is the mass fraction ofspecies i, and QLHV,i is the lower heating value of species i. The heating value of castor oil wasnot included in the fuel conversion efficiency calculation because it is assumed to act solely as alubricant, and thus would be inert in the combustion process.

Figure 2 summarizes a key result of the experiments. The figure shows an essentially monotonicincrease of power with efficiency for all of the fuel blends tested. Hence, maximum power is achievedcoincident with maximum efficiency. Note that for stable operation the efficiency is between 0.7%and 9.3% depending on fuel composition, air-fuel ratio, rotational speed, and engine load.

Figure 3 is a plot of the fuel conversion efficiency versus equivalence ratio. Note that the optimalengine operation occurs near stoichiometric conditions. For rich operation (φ > 1.2) the efficiencyand power output are very low. However, adequate efficiency and power production are achievedunder lean conditions up to the limit of φ ≈ 0.7, after which stable operation is not achievable.

The equivalence ratio, φ, is calculated by equation (12)

φ =(mmethanol + mnitromethane)/mair

(mmethanol + mnitromethane)/mair,stoich(12)

under the assumptions that air is composed of 21% O2 (by volume) and 79% N2, and the castor

7

0%

1%

2%

3%

4%

5%

6%

7%

8%

9%

10%

0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8 2.0Equivalence Ratio

Effic

ienc

y

Mix AMix BMix C

Figure 3: Plot of fuel conversion efficiency vs. equivalence ratio. The stoichiometric air-fuel ratiofor mixtures A, B, and C are 6.11, 5.43, and 4.62 respectively.

oil lubricant is an inert species. The stoichiometric air-fuel mass ratios for fuel mixtures A, B, andC are 6.11, 5.43, and 4.62 respectively.

This result is in contrast to the generally held view that model engines operate very fuel-rich.The reason for the discrepancy is that the engines do not operate as premixed systems. Becausesome of the fuel-oil mixture is used as an in-cylinder cooling mechanism, the engine exhausts alarge amount of this material unburned. The fuel-laden exhaust that gives the impression of richoperation does not result from overall rich combustion, but from an unburned wall film. The highlevels of unburned fuel in the exhaust made accurate emission measurements problematic becauseof the narrow operating range of standard gas analyzers. Improved emission measurements will beobtained in future work.

The maximum steady-state power was 82.9 W. The maximum power and efficiency were ob-tained with fuel mixture B. From the earlier discussion regarding the volumetric energy densityof nitromethane as compared to methanol, we expected maximum power to occur with maximumnitromethane content—a result that we did not get. It appears, therefore, that ignition timing(which in this engine is compression controlled) plays an important role in the power performanceas well. That is, the lower ignition temperature of nitromethane would suggest ignition earlier inthe compression stroke than would occur for a mixture with less nitromethane. In addition, thelatent heat of methanol is substantially higher than that of nitromethane, which may lead to aslightly cooler charge in the cylinder, delaying ignition. The management of ignition timing by thismechanism is difficult to control over a broad operating range.

Figure 4 is a plot of the cylinder pressure versus elapsed time for a motored cycle (no combus-

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0.0

1.0

2.0

3.0

4.0

5.0

6.0

-0.02 -0.01 0.00 0.01 0.02 0.03 0.04 0.05 0.06

Time, sec

Pres

sure

, atm

closed carburetorhalf-open carburetoropen-carburetor

Figure 4: Cylinder pressure vs. time for a motored cycle with the carburetor fully closed, half-open,and fully open.

tion) with the carburetor fully closed, half-open, and fully open. This plot captures two completerevolutions: the compression, expansion, exhaust, and intake strokes. With the carburetor fullyclosed, a large pressure drop develops during the intake stroke. This pressure drop leads to pumpinglosses. There appears to be no difference in pumping losses between the half-open and fully opencycles. The peak cylinder pressure is 5.2 atm, substantially lower than the expected value of 12.7atm.

For the motored cycle, it is assumed that the local maxima of cylinder pressure correspond toTDC, the local minima correspond to bottom dead center (BDC), and that crank degrees varylinearly with time between these extrema. The height of the piston may be found using equation(4), and thus a relationship between cylinder volume and pressure is obtained. Figure 5 is a plot oflog P versus log V for the motored cycle with the carburetor fully open. From this plot we can seethat the slope is nearly constant, thus the polytropic relation, PV n ≈ constant. Therefore, we canapply the polytropic compression relation of equation (2) to solve for the polytropic exponent of thisengine, n = 0.85. The poor compression characteristics are almost certainly the result of leakagepast the piston ring. While this leakage allows the lubrication system to be greatly simplified, theperformance of the engine suffers.

Cylinder pressure measurements during fired cycles required rotational speeds that exceeded thetime response of our pressure measurement system. We will obtain cylinder pressure measurementsfor fired engine cycles in the future.

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5

5.1

5.2

5.3

5.4

5.5

5.6

5.7

5.8

-6.3 -6.2 -6.1 -6 -5.9 -5.8 -5.7 -5.6 -5.5 -5.4 -5.3 -5.2log V

log

P

compressionexpansion

Figure 5: log P versus log V for the motored cycle with carburetor fully open. PV n ≈ constant.

Discussion

Leakage during the compression stroke does not allow the incoming mixture to reach thetemperatures necessary for autoignition. The spontaneous ignition temperatures of methanol andnitromethane are 847 K and 692 K respectively [18]. Under ideal compression (n = 1.3), thetemperature achieved after the compression stroke is 629 K. It is apparent, therefore, that whenthe engine is cold the reaction must be initiated at the platinum catalyst of the heated glow plug.Once the engine achieves its normal operating temperatures, which we measured to be between 420K and 520 K, the incoming charge is heated somewhat and the compression heating can then reachignition temperatures in the presence of the platinum catalyst without resistively heating the glowplug.

Although several modifications to the engine design would make it easier to analyze, the resultsobtained give a fairly clear picture of the dominant phenomena in small reciprocating engines.These engines appear to operate as some combination of a diesel mode and an HCCI (homogenouscharge compression ignition) mode. The HCCI component burns the fuel-air mixture preparedas the liquid fuel bleeds into the air stream through the carburetor. In addition, a wall film ofunevaporated fuel and lubricant forms on the cylinder surfaces. Fuel evaporation from this filmproduces a relatively rich mixture near the walls that might sustain a diesel-like combustion process,where the fuel vapor comes not from fuel spray but from a wall film. It is interesting that film fedcombustion occurs in this centimeter scale engine since recent results have shown that for smalldevices, a thin liquid wall film has higher surface-to-volume ratio than would droplets generated instandard ways [19, 20]. The recent papers on continuous small fuel-film combustors demonstratethis phenomenon.

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A rough estimate of the convective heat transfer from the engine assuming a transitionalReynolds number, surface temperature of 500 K, and surface area of 10 cm2 gives 10–20 W ofheat loss. While significant in comparison to the mechanical power production, this value is onlya few percent of the total chemical energy of the fuel. Incomplete combustion appears to be thedominant path of energy loss.

There is great potential to use an understanding of current model production engines to generateimproved designs for portable power applications. The most important issues that should beaddressed are the lubrication system, cooling, ignition control, and the fuel-air delivery system.

The lubrication system must be modified before these engines can be used for many portablepower generation applications because the addition of inert oil into the combustion chamber reducesthe power density of the system by lowering the reactant concentration/volume ratio and causesexcess emissions of unburned hydrocarbon oils. The addition of inert oil into the fuel tank alsoreduces the extractable energy density of the fuel (the overall energy density is higher, but theoil is not fully combusted). We excluded the heating value of the castor oil from the efficiencycalculation because we assume that it does not combust (and observe much residual oil in theexhaust stream). However, if the castor oil does contribute to the power production then it mustbe included in the efficiency calculation and the peak efficiency measured would be 5.8% ratherthan 9.3%. Furthermore, the equivalence ratio calculation must also be adjusted to include themass flow rate of the castor oil and the additional oxidizer necessary for stoichiometric reaction.

The design of the fuel-air delivery system should be improved to facilitate mixing of the fuel-air mixture prior to ignition. The carbureted design does not sufficiently atomize the reactants,resulting in a non-homogeneous mixture and leading to incomplete combustion. Part of this incom-plete combustion is intentional as a cooling source, but much better control is needed if efficiencyis to improve. Furthermore, the needle valve adjustment on the carburetor is very sensitive andrequires user intervention. For practical application the reactant delivery process should utilize afeedback controller requiring no end-user input. The O.S. Engines FS-30-S achieved maximum per-formance with the fuel mixture containing 62% methanol, 18% castor oil, and 10% nitromethane.We expected that engine performance would increase with higher concentrations of nitromethane,however that was not the case. The engine ignition timing is indirectly controlled through the fueland glow plug properties and the compression process. As with HCCI engines, suitable techniquesmust be developed in order to control the ignition process.

Acknowledgements

This work was supported by the National Science Foundation grant CTS-0212163. The assis-tance of Israel Figueroa in generating some of the preliminary data from the engine dynamometeris greatly appreciated.

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