2005-01-0246 MPI Air/Fuel Mixing for Gaseous and Liquid LPG

Paul Baker and Harry Watson University of Melbourne

Copyright © 2004 SAE International

ABSTRACT

This paper presents a parametric, experimental study of the performance of gas and liquid propane injection in a spark ignition, multi-point port injected (MPI) engine. An inline, six cylinder engine is used over a wide range of speeds and torques, and the air/fuel ratio, compression ratio and injection timing are all varied. The engine was mapped at the standard compression ratio of 9.65:1 with the original, gasoline MPI system, propane gas MPI, and single point, throttle body, propane gas injection. Gas and liquid propane MPI are then tested at a compression ratio of 11.7:1.

Contour plots of thermodynamic efficiency and the specific emissions of HC, NOx, CO2 and CO over the torque/speed range are presented and compared. The results show significant differences in performance between gas and liquid propane MPI injection, as well as the MPI and throttle body gas injection. Previous research has in part attributed this difference in performance to the increased volumetric efficiency of the liquid propane injection. This paper examines gas and liquid injection of propane and proposes that the difference in the method of mixing also significantly affects engine performance. Significant improvements in emissions and thermal efficiencies were achieved when compared with gasoline, eg. specific emissions reductions of 88% for HC, 45% for NOx, 40% for CO2, 92% for CO and a rise of 27% in thermal efficiency.

INTRODUCTION

Over the past few decades, extensive research has been undertaken with gasoline, on multi point port injection (MPI) injector nozzle design, droplet distribution, injection timing, fuel pooling and other factors relating to combustion in S.I. Research has been driven mainly by government legislation requiring ever lower emission standards and improved fuel economy. The science and art of port fuel injection has reached the stage where first injection first fire can be successfully accomplished (Honda 2002).

In the past two decades, complacency with security of oil supplies and confidence of meeting emission targets with gasoline combustion systems has seen a reduction in research effort into alternative fuels, with the result that it has slipped behind that of gasoline research. Current world events, rapid fluctuations in crude oil prices, together with an increasing sense of urgency with respect to global warming has thrown the spotlight back onto alternative practical energy sources. One line of investigation has seen the development of diesel engines, which can offer high efficiency and economy. But a major draw back with diesel combustion systems is the emission of particulates. Current research indicates that the smaller particulates, PM10 and smaller, pose the greatest risk as carcinogens(Walsh 2004).

LPG stands out as a viable alternative with a H/C ratio of 2.63:1 compared to gasoline and diesel with H/C ratios of around 1.8 to 1.85:1, offering potentially a much lower CO2 specific emission. Also, being a fuel with simple molecular structure, primarily being a mixture of propane and butanes, the fuel has a low tendency to form aromatics or particulates on combustion. See table 1 for a more comprehensive comparison of gasoline, diesel and LPG. The fuels used in this research are: (i) low octane unleaded gasoline and (ii) LPG with approximately HD5 composition (approximately 95% propane).

Table 1. Fuel Properties

Fuel Type

Res. Octane

No. (RON)

Lower Heating Value

(MJ/kg)

Lower Heating Value (MJ/l)

Rel Density

Stoic A/F

Ratio (Vol

basis)

Stoic A/F

Ratio (mass basis)

LPG (HD5)

98–103 46.33 23.63 0.51 24:1 15.7:1

Gasoline 91-93 44.2 31.82 0.74 60:1 14.7:1

Diesel _ 43.25 35.9 0.83 _ 14.5:1

The objective of this paper is to report on a comprehensive, steady state investigation of the influence of gas and liquid injection under stoichiometric and lean burn conditions in LPG-MPI and throttle body injection (TBI) systems.

Previous studies have indicated that LPG can operate under lean burn (leaner than gasoline) with improvements in exhaust emissions.(Watson 1982; Farook 1994; IMechE 1996; Shinichi G 1999) This paper includes the effects of lean burn when applied to LPG injection systems, on the basis that lean NOx catalyst technology may be applicable to engines in regions of the world where ultimate emission standards are not yet in force, but where commitment to greenhouse gas abatement makes LPG an attractive option.

Table 2. Engine specification

Bore x Stroke 92.26 X 99.31 mm

Number of Cylinders 6

Engine Configuration Vertical inline

Connecting rod length 153.85 mm

Compression Ratios 9.65:1 and 11.7:1

Spark Timing M.B.T.

Air/Fuel ratio range (λ) 1.0, 1.1, 1.2, 1.3 and 1.4

Coolant temp. 85 °C

Valve Train Single OHC, 2 valves per cylinder

In this study, the engine was an Australian Ford Falcon (2000 model) six cylinder inline, 4.0 litre spark ignition unleaded gasoline MPI engine. It was used to measure engine performance and emissions in OEM format on gasoline with a sequential MPI system, LPG TBI – gas phase, LPG MPI – gas phase and LPG – MPI liquid phase systems. Brake thermal efficiency and specific emissions were calculated from direct measurements for a wide ranging torque-speed map at steady state conditions. Parameters systematically varied for each system include air/fuel ratio, injection timing, and compression ratio.

TEST APPARATUS A summary of the test engine specification appears in table 2.

Figure 1. Engine test cell set-up.

The engine was also equipped with variable cam timing and a two-state induction runner length (broad band) system The broad band system was only used when measuring the maximum torque curve for each system. While conducting emission trials, it was set to the short runner path. For the sake of simplicity, the variable cam timing feature was not used.

The standard Ford EEC V engine control unit (ECU) was replaced by a MOTEC M48 ECU which allowed ease of manipulation of fuel injection pulse width and spark timing via a personal computer (PC) connected to the ECU via an RS232 link. Figure 1 shows a schematic of the test cell set-up.

A significant problem when measuring the mass flow rate of LPG is that the fuel is a low density (only half the density of water), saturated liquid operating under pressure. A liquid LPG system requires a circulatory system where fuel is pumped at a pressure above the tank pressure to the fuel rail. The unused fuel returns back to the tank, thus necessitating the need for two measurements (supply and return) if in-line mass flow meters are to be used. This would result in a reduction in resolution as the errors of both instruments would compound. The problem is further exacerbated by the fact that the fuel is stored in a heavy pressure vessel, so any gravimetric change in mass by the low density fuel would be masked by the large tank mass, limiting the signal resolution. The method chosen to overcome this problem was to measure gravimetrically using a beam balance mechanism where the large tank mass is cancelled by a counterbalance weight thus allowing small changes in mass due to net fuel loss, to be detected and measured. The beam balance was used in “force balance” mode by using a load cell to transmit the decreasing fuel mass to the data logging computer

which then computes the time taken for a predetermined net mass loss, thus determining fuel mass flow rate.

The mass flow rate of air was measured directly with a Hitachi model AFH80-X21 hot wire anemometer located at the air intake.

The air/fuel ratio was measured via three different methods: • Computed from the mass flow readings of air and

fuel, • via a Bosch LSU 5-wire wide band heated exhaust

gas oxygen (HEGO) sensor, and • via an NDIR five-gas exhaust gas analyser.

In this way, the measurements could be crossed checked for accuracy and an average lambda value computed. Fuel rail pressure and temperature (liquid and gas) were also measured and logged to determine the state of the fuel upstream of injectors. Fuel rail pressures and injection timing were varied for both gas and liquid injection systems to find the setting that gave the optimum operating performance and lowest emissions, using HC emissions and imep as indicators.

An eddy current type Heenan and Froude Dynamatic II brake dynamometer was used in 1st quadrant speed control mode. The relative humidity was calculated by the data logging computer via wet and dry bulb temperature signals from a forced draught psychrometer and barometric pressure was entered via the keyboard, allowing brake torque and power corrections.

Keihin type 3 gas injectors were used for the MPI-Gas application, replacing the original gasoline injectors, and two Keihin type 2 gas injectors were used in the TBI-Gas application, located just upstream of the throttle plate. Seimens bottom feed injectors were used for the liquid phase MPI application. Fuel rail pressure for the MPI gas system was maintained to give a 375 kPa pressure drop across the injector . The fuel rail pressure was modulated by a MAP feedback signal from the manifold. Fuel rail pressure for the liquid injection system was achieved by the application of a pressure regulated non-return valve located in the return section of the fuel rail. Rail pressure for the liquid injection system did not appear to be quite as critical with a pressure of 10 bar maintained upstream of injectors.

EXPERIMENTAL PROCEDURE

The strategy employed was to standardise all settings, except the variable being tested; i.e. all tests were carried out at minimum spark advance for maximum brake torque (MBT), the broad band manifold setting was set to short runner and the VCT disabled.

The parameters varied were fuel system/fuel type, lambda and compression ratio. Only one parameter was

varied at any one time while all other parameters were fixed. At each setting, the engine was tested for mapping points ranging from 1000 to 4000 rev/min in increments of 500 rev/min, at torques from no load to wide open throttle (WOT), in increments of 50 Nm. Fuel systems/types, tested at the standard compression ratio of 9.65:1 were: (i) standard OEM gasoline MPI system with gasoline (ii) propane (approximately HD5 blend) in gaseous

phase with throttle body injection (TBI), and (iii) propane in gaseous phase with MPI.

Systems tested at the higher compression ratios of 11.7:1 were propane MPI in gaseous and liquid phase injection modes. All systems were tested at lambda values of 1.0, 1.1, 1.2, 1.3 and 1.4.

After installing each fuel injection system, the injection timing was optimised over the speed and load range, to minimise HC emissions and coefficient of variation of indicated mean effective pressure (CoVimep), while maximising CO2 (as an indicator of combustion efficiency).

MIXING CHARACTERISTICS

In the mid 90’s, research was beginning to emerge indicating that the gas mixing processes in CNG and LPG automotive gas systems may be more of a challenge than first thought.(Abraham J. 1994; Abraham, Minnesota et al. 1995; Das 1995) 3-D Modeling carried out by Abraham comparing the mixing of a methane gas jet from an injector with a liquid spray injection as used in direct injection in diesel engines, demonstrated that the liquid injection gave faster mixing. Das, in his work with CNG MPI systems on a 6 cylinder SI engine discovered that the location of the injector in the runner, had a significant effect on combustion efficiency and engine out emissions. The closer to the valve, the. injector was located, the worse the combustion process. He concluded that this was very likely due to poor mixing of the gas jet co-flowing with the manifold air. Inspection of figures 2 and 3 reveals that the overall values of brake thermal efficiency for the LPG MPI gas system are slightly lower than for the gasoline MPI system, and that there is a bias of the brake thermal efficiency peak to opposite sides of the graph for each fuel type. i.e.. gasoline liquid spray injection shows a strong bias to the lower speed / high torque section of the map, while LPG MPI gas injection shows a bias to the high speed / high torque side of the graph.

It has been proposed that cylinder-to cylinder variations in A/F ratio could be different amongst the fuelling systems and thus contribute to the shift in the shape of the efficiency maps. This research has indicated however, that a mal-distribution of A/F ratio across

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Figure 2(a). Brake thermal efficiencies of gasoline with standard MPI system at compression ratio 9.65:1 for Lambda 1.0.

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Figure 2(b). Brake thermal efficiencies of gasoline with standard MPI system at compression ratio 9.65:1 for Lambda 1.2.

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Figure 2(c). Brake thermal efficiencies of gasoline with standard MPI system at compression ratio 9.65:1 for Lambda 1.4.

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Figure 3(a). Brake thermal efficiencies of LPG with MPI gas system at compression ratio 9.65:1, Lambda 1.0

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Figure 3(b). Brake thermal efficiencies of LPG with MPI gas system at compression ratio 9.65:1, Lambda 1.2

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Figure 3(c). Brake thermal efficiencies of LPG with MPI system at compression ratio 9.65:1, Lambda 1.3

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Figure 4(a). Brake thermal efficiencies of LPG with TBI system at compression ratio 9.65:1, Lambda 1.0

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Figure 4(b). Brake thermal efficiencies of LPG with TBI system at compression ratio 9.65:1, Lambda 1.2

Figure 4(c). Brake thermal efficiencies of LPG with TBI system at compression ratio 9.65:1, Lambda 1.3

Figure 5 Cylinder-to-cylinder variation in A/F amongst the fuelling systems at 1500 r/min, 80Nm & Lambda 1.0 (nominal)

cylinders occurs independently of the type of fuel system used, TBI or MPI. A typical example of A/F distribution is shown in figure 5 for three fuelling systems at a moderate speed-load point. The trend for leaner mixtures in cylinders 1 and 6 is evidenced across the fuels and supply methods (Baker 2004), however, although the variations are slightly more pronounced with the higher compression ratio variants of the engine. Although these variations influence the engine’s efficiency at high values of lambda when some cylinders approach the lean limit, these variations are likely to have common effects amongst the fuel supply and injector location positions.

It is proposed that the phenomenon of differing efficiency map shapes is primarily due to the fundamental difference in mixing characteristics between fluids of the same phase, compared with that of different phases.

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For fluids of the same phase, the mixing process is primarily reliant on turbulence within the media to affect mixing. As Abraham’s model shows(Abraham J. 1994), turbulence at the periphery of the jet caused by the transfer of momentum of the gas jet to the adjacent air creates the shear necessary to promote mixing only at the interface of the gas jet with the surrounding air. This mixing process may be enhanced by additional shear caused by bulk turbulence within the air stream. The mixing mechanism of fuel sprays is quite different. The very fine liquid droplets within the liquid spray, have a much greater density than the surrounding air, allowing greater penetration and greater momentum transfer to the air creating local shear in their wake. Vaporization of the droplets occurs as they travel through the air medium, leaving a vapor trail behind them. This mechanism greatly assists the bulk distribution of vapor throughout the air. The greater interface surface area created allows greater diffusion rates and faster mixing.

The shape of the brake thermal efficiency contour plots in figure 2 are fairly typical for gasoline, where the peak

efficiency is shifted left due to mechanical friction losses. Friction power is a function of engine speed, so the lower the speed, the lower, the friction losses. Hence, at lower speeds, the net useful power output will be proportionally greater due to lower friction loss. It is proposed that the peak occurs at the high torque region of the map not only due to the higher proportional net useful power output as described above, but also to higher turbulence within the combustion chamber, and within the intake runners due to the higher flow rates. The higher turbulence within the combustion chamber will promote a faster burning flame, and it will also aid the final mixing process of the reactants prior to combustion thus promoting greater combustion efficiency. In the same way, higher flow rates will promote higher turbulence within the intake system thus assisting the mixing process.

The shift of the thermal efficiency peak to the right in figure 3 may be explained by the higher flow rates and hence, higher turbulence at the higher rev range, required to promote adequate mixing of the gas MPI system. Poor mixing may also explain why the brake thermal efficiency figures are slightly lower for the LP gas MPI system. It may also be noted that the maximum torque curves for all lambda values for the gas MPI fuel system are lower than those for the gasoline liquid MPI system. (It must be emphasized here that the engine is capable of producing much higher toques than those shown in this study. The primary purpose of this study is to highlight the mixing processes so the broad band intake system and VCT were not used.) It has been well documented that a loss of power can be expected with gas mixing systems due to a lower volumetric efficiency caused by the gas displacement of air in the intake charge and the lack of cooling of the charge due to enthalpy absorption as a result of vaporization of the fuel.

It could be argued that the differences between figures 2 & 3 could be attributed to the combustion characteristics of the two different fuel types, and not to the fundamental differences in mixing. A study of figures 3 and 4 can shed light on this question. The only difference between the fueling systems used in figures 3 and 4 is the location of the injectors. All parameters have been kept constant except for the location of injection. Both systems use LPG gas injection, but the system used in figure 4, replaces the six Keihin Type 3 injectors located at each port, (same location as the conventional gasoline injectors) with two larger flow rate Keihin Type-2 gas injectors, located just upstream of the throttle body, that is, the same location as the gas mixer in a conventional second generation type LP gas mixer system. Examination of figure 4 shows that the peak of the thermal efficiency has shifted to the left, into the lower rev range, similar to the gasoline graphs. It also shows a slight increase in brake thermal efficiency above gasoline but drops away at lambda 1.3. Figure 4 also shows that the maximum torque curve is lower for TBI – gas.

The shift of peak brake thermal efficiency to the left supports the proposition that gas mixing is a strong influence on the shape of the brake thermal efficiency map. A computational Fluid Dynamic (CFD) model developed for the intake system in question by Liew,(Liew 2004) shows that significant levels of turbulent mixing occur within the intake system for a TBI system. The model predicts that even with the additional turbulent mixing occurring over a longer inlet runner path, imperfect mixing still results due to the inherent difficulty of gas mixing. The model also predicts uneven A/F distribution from cylinder to cylinder. Results from this investigation, as shown in figure 5, confirm this. The longer residence time in the inlet system also has some down side. The larger volume of gas/air mixture in the inlet system increases the probability of backfire. During the trials, at certain operating points, backfire occurred quite frequently, whereas with the MPI system, backfire did not occur. Figure 4 also shows that maximum torque deteriorates at higher engine speeds. It is thought that this could be due to a lower volumetric efficiency introduced by the TBI system. Again it could be argued that factors other than fuel mixing, have been dominant in shaping the thermal efficiency maps for the TBI and MPI gas systems. Figures 6 and 7 show brake thermal efficiencies for LPG MPI systems at an engine compression ratio of 11.7:1.

The only variation between these two sets of figures is the method of injection. Figure 6 represents LPG MPI gas injection and figure 7, LPG MPI liquid injection. For both systems, there appears to be a shift to the left with the higher compression ratio, but there is still a significant difference in the bias within the graphs. It is interesting to note that with the gas injection system, there is a dominant peak on both sides of the map, evident at stoichiometric and at lambda 1.4. It is thought that the higher compression ratio may be assisting the in-cylinder mixing/combustion processes in the lower rev range. It is also pointed out that the brake thermal efficiencies achieved for the higher compression ratio for LPG, especially liquid phase, are considerably higher than for gasoline at the standard compression ratio. Thus the higher knock rating of LPG can be used to great advantage. This is important as it means that LPG can compete with gasoline economy. On a mass basis, LPG wins hands down with fuel efficiency, but on a volume basis, gasoline is ahead due to the lower density of LPG. Fuel is currently purchased per unit volume, thus giving gasoline the edge. With the higher efficiencies reported in this paper, it is calculated that the specific fuel consumption, volume basis for LPG, especially LPG liquid phase, rivals that of gasoline. For comparison, in figure 8, the brake thermal efficiency of the standard MPI gasoline system, over the lambda range of 1.0 to 1.4, have been plotted against those of all LPG systems tested. The values have been normalised by dividing each value by the corresponding value of gasoline at stoichiometric.

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Figure 6(a) Brake thermal efficiency for LPG MPI Gas phase at lambda 1.0, compression ratio 11.7:1

Figure 6(b) Brake thermal efficiency for LPG MPI Gas phase at lambda 1.1, compression ratio 11.7:1

Figure 6(c) Brake thermal efficiency for LPG MPI Gas phase at lambda 1.4, compression ratio 11.7:1

Figure 7(a) Brake thermal efficiency for LPG MPI Liquid phase at lambda 1.0, compression ratio 11.7:1

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Figure 7(c) Brake thermal efficiency for LPG MPI Liquid phase at lambda 1.4, compression ratio 11.7:1

Figure 8 Normalised brake thermal efficiencies of fuels systems tested.

The brake thermal efficiency trends in figure 8 show that there was an improvement in thermal efficiency for all LPG fuel systems except MPI-gas at the standard compression ratio. Liquid phase MPI at compression ratio 11.7:1 returned the best results with a consistent 25% increase over gasoline at stoichiometric, from lambda 1.2 through to 1.4.

EMISSIONS

The other important part of this investigation is to study the comparison of emissions from gasoline MPI, and LPG MPI gas and liquid systems, and their behaviour under lean burn conditions. For brevity, the results have been condensed into a series of graphs showing the comparison of engine out brake specific emissions of HC, NOx, CO and CO2. For the purpose of comparison, the results have been normalised by dividing each value by the corresponding stoichiometric value for gasoline and averaging them over the entire map range for each A/F setting. The results are shown in figures 9 through 12.

Figure 9 Normalised brake specific emission for HC.

Figure 9 shows that HC emissions for the LPG fuel systems, in general, are almost an order of magnitude below those of gasoline at stoichiometric. They remain fairly constant out to lambda 1.2 but then begin to rise. It is noted that both MPI-gas systems, irrespective of compression ratio, follow the same curve from lambda 1.0 to 1.2 with slightly higher emission figures than TBI and MPI-Liquid phase. It is also noted that both TBI at standard compression ratio and MPI-Liquid phase at 11.7:1 compression ratio follow the same curve out as far as 1.3 before TBI suffers deterioration in performance. As stated above, both TBI-gas and MPI-liquid phase display similar thermal efficiency maps indicating better mixing than MPI-gas phase systems. The results reinforce the gas mixing hypothesis.

Figure 10 Normalized brake specific emissions for NOx

Given that NOx formation is a function of combustion temperature, at first glance it would be expected that TBI should give the highest results, as the initial charge temperature should be higher than either liquid injection or even MPI gas injection. Figure 10 shows NOx production for LPG MPI–gas at standard compression ratio giving the highest results, peaking at around Lambda 1.1. A careful inspection of figures 3 and 4 will show that maximum torque is consistently higher for MPI-gas than for the TBI system. This would imply higher combustion temperatures for the MPI-gas case, thus explaining the higher NOx value for the MPI-gas system. All LPG systems except for the MPI-gas system operating at standard compression ratio, show lower NOx values than for gasoline. In the case of MPI-liquid phase at 11.7:1 compression ratio, the NOx output is only 55% that of gasoline at stoichiometric. The torque curve for the MPI-liquid phase system is actually higher than that of gasoline, so it could be reasonably expected that the NOx value should also be higher. In fact, it is considerably lower. One possible explanation for this is that at the higher compression ratio, the chamber surface area to volume ratio is higher, thus promoting more effective cooling of the in-cylinder combustion gases. The down side of this effect is that it also means that the quench layer volume will be a higher proportion of the clearance volume, thus producing higher HC and CO emissions. Figure 9 shows that there is little

evidence of high HC for the high compression trials, either MPI-gas or MPI-liquid. Close inspection of figure 12 shows that at stoichiometric, there is marginally higher CO emission for the high compression trials compared to TBI at the standard compression ratio.

Figure 11 shows the normalized specific emissions for CO2. Again, there is a striking difference between gasoline and LPG. i.e. Emissions for the high compression trials producing only 64% of the value to that of gasoline at stoichiometric.

Figure 11 Normalized brake specific emissions for CO2

The figures for the LPG fuel systems tend to differentiate more between compression ratio’s than fuel system types. MPI-gas and MPI-liquid for the high compression trials producing almost identical curves and MPI-gas producing slightly higher emissions than TBI at the same compression ratio reflecting a lower combustion efficiency. As already mentioned, the curves shown in figure 12 for CO emissions, show a slight increase at stoichiometric for the high compression trials compared to TBI at the standard compression ratio. All LPG trials with the exception of MPI-gas at standard compression ratio, produced a CO value of only around 50% that of gasoline. The striking exception is MPI-gas at standard compression ratio, which shows a value at stoichiometric of 300%, dropping down to 40% of stoichiometric at lambda 1.1. These figures are consistent with our mixing hypothesis, where it is predicted that MPI-gas at standard compression ratio should produce higher levels of CO due to poorer combustion resulting from poor mixture preparation compared to the liquid injection systems and TBI.

Figure 12 Normalize brake specific emissions for CO

There is a slight increase in CO emission level at stoichiometric for MPI-gas high compression compared to MPI-liquid phase high compression, also consistent with poorer mixing with the gas system.

SIMULATED DRIVE CYCLE EMISSIONS

Methodology

The foregoing presentation represents a comparison of the ‘whole-of-engine map’ values. Another alternative is to focus on frequently used torque and speed points in urban driving. Such points are often called ‘World-wide mapping points’. Appendices B and C present results for thermal efficiency and exhaust emissions at the 200 kPa, 2000 revs/min mapping point. In the following we focus on the part of the torque speed map that is used in actual driving rather than the entire map average values or a single point in the map. In vehicle emissions and fuel consumption measurements the car is driven over a prescribed drive cycle. In so doing the engine follows a trajectory on the engine map. Such a trajectory for a range of driving cycles can be seen in figure 13. (Watson 1999)

It is possible to convert these trajectory diagrams into particular engine frequency map for a given vehicle as shown in figure 14. As the Euro2/3 drive cycle is made up essentially of constant speed or constant acceleration/deceleration maneuvers it is the cycle for which predictions from engine maps can be expected to be related to vehicle measurements.

Figure 13 Torque-speed trajectories for a 1.625 tonne Ford Falcon car with the engine tested here and an automatic transmission.

Figure 14 Frequency map for the Euro2/3 drive cycle.

The values at each torque and speed point in the engine map (converted to grams per second) are multiplied by the fractional frequencies in figure 14 and summed to give the drive cycle average value, in this case for the Euro 3 cycle. Of course these results are for hot start

operation only engine-out emissions. Ignoring the cold start effects will result in the relativities between gasoline and LP gas, likely to be in favor of gasoline for HC, CO and thermal efficiency since the cold wall fuel wetting issues in the ports and manifold are not so significant for a gaseous fuel.

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-50

0

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0.80.8

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1.52.5

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m)

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EURO 2/3 TIME FREQUENCY DISTRIBUTIONFOR FORD AU ENGINE

Figure 15 Euro 3 drive cycle weighted emissions of HC, CO and NOx across the range of LP gas fuel supply technology compared with the reference gasoline (ULP)

Figure 15 shows the simulated Euro3 drive cycle emission results. In general the trends are similar to the

whole-of-map values and the explanations for the differences hold. The effect of cycle weighting influences the magnitude of the trends. For HC, all the LPG results are much less than the base engine gasoline data. At stoichiometric about 15% of gasoline, and even under lean burn conditions 30% less than stoichiometric gasoline, with the increased compression ratio reducing the HC emissions in spite of the increased combustion chamber surface-area to volume ratio.

The CO trends are similar to those for HC. Emissions reductions from LPG use are in the range of 50 to 90% lower than those for gasoline.

The NOx trends show reductions below those for gasoline with the higher compression ratio results generally giving higher NOx as might be expected. At the high compression ratio the liquid phase shows an advantage over the gas phase MPI presumably because of lower mixture temperatures that result from the fuel evaporation. It is noted that at the leanest mixture the engine out emissions are about half of the present Euro3 diesel emission standard and reduced by 75% from stoichiometric gasoline operation.

Figure 16 Simulated Euro 3 Drive Cycle Emissions CO2 and Thermal Efficiency across the range of LP gas fuel supply technology compared with the reference gasoline (ULP)

Figure 16 comprises the CO2 and fuel consumption results. As the CO2 is always the major contributor to greenhouse gas emissions from transport emissions, the benefits in this emission reduction is a significant argument for the application of this fuel. The reduction in CO2 emission for liquid LP gas injection is 22% at stoichiometric increasing to 28% for lean burn, with the gaseous injection a few per cent less. Although this reduction is a combination of the increased compression ratio and lean burn effects the benefits are possible because of the higher octane number of the gaseous fuel.

The thermal efficiency results in figure 16 exemplify the real increases in efficiency possible with gas technology over gasoline. Experience with bi-fuel cars in comprehensive vehicle tests across a range of LP gases(Watson 2000) and the 9.65:1 CR results here demonstrate that ordinarily both fuels deliver fuel consumption on an energy basis within a few per cent. With the increase in compression ratio and choice of liquid fuelled LPG at lambda = 1.4, the improvement in thermal efficiency is 27% relative to low compression ratio stoichiometric.

DISCUSSION The thermal efficiency results above can be summarised as follows: The change from gasoline to LPG produced

marginal changes in efficiency except for leaner mixtures where the slightly faster burning rates of LPG are likely to show their benefits.

The increase in compression ratio by two numbers to 11.7 (an increase to 12.9 was also tested but not comprehensively Baker (Baker 2004)) might expect to lead to an improvement in thermal efficiency of about five percent if the efficiency follows the Otto cycle increase. It can be seen in Appendix A that the increase in compression ratio was achieved by significant piston and some cylinder head design changes to achieve a canted compact hemispherical chamber with carefully developed squish regions and squish divergent angles. In an earlier program a Ford Falcon AU version of the engine converted to natural gas and 15.7:1 compression ratio achieved a peak thermal efficiency of nearly 41% (Das 1995). Production pistons from this engine series were modified to reduce compression ratio and to fit within the reduced block height of the BA version of the Falcon engine used in the present work.. Thus the increase in efficiency from the increased compression ratio comes from a substantial and well proven chamber design. Not only does this chamber confer efficiency improvements of 5 to 7% because of the faster burn combustion but it also results in an increase in peak torque to 370 Nm from 354Nm for the 4 L I-6 engine (Das 1995; Baker 2004).

The shift to lean mixtures of lambda 1.4 results in about three times the improvement in efficiency

greater than that for gasoline. Part of the gain is the result of the reduced pumping losses with lean mixture shift (there is no EGR used in this engine). Again the faster flame speed, enhanced by the combustion chamber design contributes to the achievement of around 10% extra efficiency.

Most surprising is the increase from liquid phase injection. The evaporation and better distribution of the fuel in the inlet port (spray versus co-flowing streams) can make some contribution. The lower mixture temperatures are evidenced by the lower NOx. The lower combustion temperatures will contribute to lower peak cycle temperatures (even though the burn is still fast as the MBT spark advance requirements differ little) which in turn will give less dissociation of CO2 and water. There is a hint of this trend in the reduced CO levels in figure 15 for example. Nonetheless the authors were surprised at the magnitude of the about 10% additional benefit over gaseous phase injection across the mixture range. It should be stressed that the same direct mass measurement of fuel consumption was used as the reference in all tests and cross checks were possible from the air flow meter/air fuel ratio measurements by both lambda sensor and exhaust analysers.

Finally we note that the 28% improvement in best achieved lean-burn LPG efficiency over gasoline stoichiometric operation is almost enough to compensate for the density deficiency of LPG. That is, a so configured LPG car might equal its gasoline counterparts miles per gallon on a liquid fuel basis.

CONCLUSION

The study looked at the results of A/F mixing characteristics of co-flowing gases, as applied to LP TBI-gas and MPI-gas fuel systems, and fuel spray injection as applied to MPI-liquid systems. The effect that mixing characteristics had on LPG lean burn regimes was also studied, along with increasing compression ratio to observe the effect on A/F mixing and lean burn. A conventional six cylinder SI gasoline engine was used to test LP TBI-gas and MPI-gas systems at standard compression of 9.65:1 and MPI-gas and MPI-liquid systems were tested at a compression ratio of 11.7:1. All tests were carried out at steady state conditions. The findings are summarised:

1. A link has been established between fuelling system and brake thermal efficiency and emissions for this engine.

2. The brake thermal efficiency maps and emission trends results suggest that fuel spray injection systems are faster and more efficient than gas phase injection.

3. Engine out emission trends indicate that there is very little difference in sensitivity to A/F ratio in the lean burn region between gas and liquid injection systems. Increasing the compression ratio appears to have a greater impact on lean burn characteristics. Increasing the compression ratio to

11.7:1 enhanced the emission outputs, increased efficiency and tolerated leaner A/F ratios. Figures achieved include reductions in engine out brake specific emissions of 88% for HC at lambda 1.0, 45% at lambda 1.0, 40% for CO2 at lambda 1.2 out to 1.4 and 92% for CO at lambda 1.1 out to 1.4. An increase in brake thermal efficiency of 26% was also achieved at lambda 1.4.

4. Going lean of stoichiometric improves emission outputs and increases brake thermal efficiency. NOx brake specific emissions was 27% that of gasoline at stoichiometric.

From a practical stand point, LPG MPI-liquid and gas systems coupled with increased compression ratio deliver superior emissions figures, power and comparable economy (volume basis) to gasoline. In the future world economic climate, as oil stocks dwindle and with greater concern on greenhouse gas emissions, LPG is well suited as an alternative fuel to gasoline, and still has a great deal of untapped potential. This study has only looked at steady state engine operation, but it is anticipated that MPI gas and liquid systems will perform better under cold start and dynamic conditions.

ACKNOWLEDGMENTS

The authors would like to express thanks to Ford Australia for the donation of the test engine, to Dockland Science Park P/L and Nick De Vries for assistance with equipment when needed.

To the late Eric Milkins, who provided guidance in several phases of the project, to whom we dedicate this paper.

REFERENCES

1. Abraham, J., U. o. Minnesota, et al. (1995). "Effects

of combustion on in-cylinder mixing of gaseous and liquid jets." SAE Paper No. 950467.

2. Abraham J., M. V., Macinnes J., Bracco F.V. (1994). Gas Versus Spray Injection: Which Mixes Faster? SAE Paper No. 940895.

3. Baker, P. A. (2004). LPG: A Comparison of Multipoint Liquid and Gaseous Phase Injection - Current PhD Thesis. Dept of Mechanical and Manufacturing Engineering. Melbourne, University of Melbourne.

4. Das, A. (1995). Optimisation of a natural gas spark ignition engine - PhD Thesis. Department of Mechanical and Manufacturing Engineering. Melbourne, University of Melbourne.

5. IMechE. (1996). Lean Burn Combustion Engines. London, IMechE Seminar Publications Ltd.

6. Farook, Y. (1994). Limits of performance of an SI engine operating on LPG and petrol - M.Eng Science Thesis. Dept of Mech & Manufacturing Engineering,, University of Melbourne.

7. Liew, G. C. S. (2004). LPG Mixing in an Engine Inlet Manifold with Throttle Body Injection - M.Eng

Science Thesis. Mechanical and Manufacturing Engineering. Melbourne Vic., University of Melbourne.

8. Shinichi Goto. (1999). Performance and Emissions of an LPG Lean-Burn Engine for Heavy Duty Vehicles. SAE Paper no. 1999-01-1513.

9. Walsh, M. (2004). Global trends in motor vehicle pollution control - 2004 Update. FISITA Paper No. F2004V023

10. Watson, D. H. C., Milkins E.E. (1982). "Comparison and optimization of emission efficiency and power of five automotive fuels in one engine." Int. J. of Vehicle Design, 3(4): 463 - 476.

11. Watson, H. C. (1999). Engine and Environmental Impacts - Euro 2 Compared with US City FTP. SAE - A Paper No. 99099.

12. Watson, H. C., Gowdie D.R.R. (2000). "The systematic evaluation of twelve LP gas fuels for emissions and fuel consumption." SAE Paper number 2000-01-1867.

CONTACT

Paul Baker – Currently undertaking a PhD Thesis in the Department of Mechanical and Manufacturing Engineering, University of Melbourne, Parkville, Vic. Australia 3010 E-mail: p_a_baker@optusnet.com.au

APPENDIX A

COMBUSTION CHAMBER DETAILS FOR STANDARD AND HI COMPRESSION ENGINES.

Standard Compression Ratio 9.65:1

APPENDIX B

Hi Compression Ratio 11.7:1

APPENDIX C