Applied Thermal Eng i neering 48 (2 0 1 2) 3 2 e 40

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Applied Thermal Engineering

journal h o m ep age: www.elsev i er .com/locate/apthermen g

An experimental and analytical investigation of a multi-fuel stepped piston engine

Peter R. Hooper*, Tarik Al-Shemmeri, Michael J. Goodwin

Faculty of Computing, Engineering and Technology, Staffordshire University, Beaconside, Stafford ST18 0AD, UK

a r t i c l e i n f o

Article history:

Received 7 December 2011

Accepted 13 April 2012

Available online 21 April 2012

Keywords:

Engine modelling

WAVE CFD

internal combustion engine

stepped piston engine

externally scavenged two-stroke engine

unmanned aircraft engine

single fuel policy

a b s t r a c t

This paper presents results of computational modelling of a stepped piston engine using one dimensional CFD code. The analysis builds upon the experimental work performed on a four-cylinder stepped piston engine for Unmanned Air-Vehicle (UAV) application. A range of variables in terms of fuels, fuelling methods and core engine parameters have been modelled and compared with actual test data. The maximum power recorded from experimental testing was 30.47 kW at 5250 RPM using kerosene JET A-1. The correlation between theoretical and experimental data is in general agreement within the bounds of the uncertainties of experimental errors and the assumptions within the numerical models. Simulation has allowed an assessment of potential for direct injection fuelling predicting minimum specific fuel consumption (SFC) of 0.273 g/kWh using indolene and 0.310 kg/kWh using simulated JET A-1.

2012 Elsevier Ltd. All rights reserved.

1. Introduction

The need for low mass engines capable of operating efficiently on low volatility fuels has been a NATO objective for quite some time [1]. The requirement remains to this day largely unfulfilled [2] except for a number of small volume exploratory applications for unmanned air-vehicle (UAV) systems. The high volatility of gasoline and therefore its deployment in theatre and onboard Navy vessels poses a very hazardous safety risk. Furthermore due to the fact that the majority of military aircraft and ground based vehicles operate on kerosene based (AVTUR, NATO F34, JP5, JP8) and diesel fuels, the logistical supply and support challenges of moving and storing AVGAS and AVGAS/oil mixed fuels for conventional two-stroke engine powerplants presents significant problems.

UAVs by their very nature have to meet very stringent targets in order to achieve low overall vehicle mass objectives. All of the onboard systems, notably air-frame, avionics, electronic surveil- lance payload, propulsion system and fuel payload must be designed to meet these minimum mass requirements. The power- plant normally represents one of the highest mass assemblies within the overall UAV system. In the case of the propulsion system, the most efficient method of combustion of low volatility fuels is achieved using compression ignition. Thermal efficiencies in excess

* Corresponding author. Tel.: þ44 (0)1785 353514; fax: þ44 (0)1785 353520.

E-mail address: p. r .hooper@staffs.ac.uk ( P.R. Hooper).

of 50% [3] have been demonstrated using two-stroke cycle diesel engines. Unfortunately these are high mass engines achieving very high power levels but at a power:mass ratio of only around0.035 kW/kg. Automotive four-stroke diesel engines are capable of achieving typical levels of 0.8 kW/kg. Unfortunately these also present too high a mass penalty for UAV application. For medium and short range applications turbine powerplants are largely unsuitable. For fixed wing UAVs the minimum cruise speed or loitering speed of small turbine powerplants is normally too high for practical surveillance missions unless rotary wing UAVs are adopted [4].

A means of overcoming this problem was identified via exper- imentation [5] with low mass spark ignition engine combustion of kerosene JET A-1. Successful operation of such a system based upon an engine not required to sustain the high pressure required for compression ignition, offers a means to meet this conflicting need. The two-stroke cycle engine has been the subject of research and development for low emission automotive applications via results observed using direct injection. Notable examples demonstrated by Schlunke [6], Duret [7] and most recently by Turner et al. [8]. The highest thermal efficiency levels with a two-stroke cycle engine are therefore likely to be achieved with direct injection of heavy fuels.

During the experimental phases of the research forming the subject of this study a range of tests were conducted to assess the effects on the performance of a stepped piston engine designed and developed at Bernard Hooper Engineering Ltd (BHE) under Ministry of Defence contract. The work included

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P.R. Hooper et al. / Applied Thermal Engineering 48 (2012) 32e40 33

ch Cm (1)

f

exploration of the effects of compression ratio variation and a range of fuel preparation methods. The results of the subse- quent computational modelling of the engine are also pre- sented, together with discussion of modelled and actual results correlation. The use of Ricardo WAVE engine simulation soft- ware has allowed investigation of the potential for direct

Table 1

SPV580LC performance to date.

Cylinders 4

Cylinder configuration 90 V-4

Bore stroke mm 62 48

Swept volume cm3

580

Dimensions including generator (L W H) mm 360 364 369injection.

2. Stepped piston engines

Power 95 RON gasoline (Stub exhaust)

Min propeller cruise load SFC

Power 95 RON gasoline (Advanced exhaust)

Min propeller cruise load SFC

kW

kg/kWh

kW

kg/kWh

30.9

0.318

35.4

0.286

The method of operation of the stepped piston engine has been presented in previous papers by Hooper et al. [5,9e14], however the operating principle is reproduced in Fig. 1.

The SPV580 engine operates with two banks of paired cross- charging cylinders. Air or air and fuel mixture is drawn into the pumping annulus, through a reed valve. As the piston ascends the charge is transferred through a crossover system to the working cylinder. Thereon the engine operates as a normal two-cycle engine, except that the crankcase isolation enables fundamental durability and operational advantages.

2.1. Stepped piston engine advantages

The crankcase isolation afforded by the stepped piston allows the following key advantages:-

Low piston thermal loading e due to positive cooling methods Re-circulatory wet sump lubrication Plain bearings No valve gear High durability with low emissions Compact low mass design Extended oil change periods Low manufacturing costs Ability to operate on wide range of fuels (no added oil) Simple compression ratio variation

The SPV580 uses a 90 V4 cylinder configuration with

a capacity of 580 cm3 (power cylinder swept volume). Air-cooled and liquid-cooled versions of the engine have been developed.The liquid-cooled engine was used for the experimental work reported in this paper. Details of the engine are presented in Table 1.

Power kerosene JET A-1 kW 30.5

Cooling system Liquid (air option)

Mass kg 17.45 (17.22)

3. Computational engine modelling

Models replicating the SPV580 were constructed with basic structures shown in Fig. 2. The external CFD junctions highlighted are configured to allow transfer of mass, flow velocity and further thermodynamic properties of the output from the parent model to the child model input. The pumping cylinder and intake/exhaust geometry was defined as the parent model while the power cylinder geometry, crossover system, transfer and exhaust ports were defined as the child model. The output from the pumping cylinder (parent model) therefore providing the starting conditions for the relevant power cylinder (child model).

3.1. Model theory

The models developed use in-cylinder heat transfer assump- tions based upon the Woschni [15] model. The expression for heat transfer coefficient within the engine cylinder was derived by Woschni based upon engine experiments with normal operation (ignition applied) and with the engine being motored externally (i.e. not firing). The heat transfer coefficient derived by Woschni is determined from the following relationship:-

h ¼ 0:0128 D 0:2 P0:8 T 0:53 v0:8

Woschni found that when the engine was motored the charac- teristic gas velocity vch, can be defined as the mean velocity of the

piston. When the engine was operated under its own combustion the velocity varies proportionally as a function of the increased cylinder pressure. The original 1967 Woschni model was employed in all models created during this programme of work. The relative heat transfer area scaling factor, Cm, in equation (1)

is related to cylinder bore area. Cm values derived were 1.016

and 1.267 for piston and combustion chamber surfaces respectively.

The Colburn analogy defined by Bird et al. [16] is used to determine heat transfer in circular ducts. Colburn’s analogy for heat transfer coefficient is defined as:-

C 2h ¼

2 rUcp Pr

3 (2)

Fig. 1. Stepped piston crossover system (SPX) e Simplified schematic (image

courtesy of Bernard Hooper Engineering Ltd).

The friction coefficient, Cf, is calculated based upon the laminar or turbulent flow regime occurring within the model.

Engine friction is allowed for using the Chen and Flynn [17] correlation. The use of Morse tests during dynamometer testing provides a means of assessing mechanical efficiency and hence friction mean effective pressure (FMEP). It is possible to define a representative friction profile for the engine CFD model. Data from Morse testing recorded at BHE has been used to determine FMEP values at various engine speeds, thereby attempting to ach- ieve closer correlation between modelled and actual test data. It should be noted that the original intention of the Morse tests was to

Fig. 2. Models of pumping (upper) and power cylinders (lower) e SPV580.

make an assessment of the engine mechanical efficiency and in particular each cylinder’s contribution to the overall power output rather than to directly assess FMEP so only a limited range of engine speeds were assessed during the tests. The prior existence of this data has proven useful for subsequent development of correlation factors within the models. The relevant equation for FMEP derived for the model based on Chen and Flynn correlation was as follows:-

experimental test data for the Wiebe model exponents. A further method for improving correlation is to derive the heat release profile from the instantaneous cylinder pressure throughout the cycle. WAVE allows input of measured cylinder pressure profiles in order to influence the heat release model. Unfortunately experi- mental test programmes during development of the SPV580 engine did not require installation of cylinder pressure transducers and sono data exists. However some exploratory work to try to improve

FMEP ¼ 0:1167 þ 0:005ðPmax Þ þ 552:764 S

RPM 2

correlation has been conducted. Aziz et al. [23] built a 125 cm3

stepped piston engine and instrumented the engine to analyse

þ 2:8689S

2

RPM 2

(3)cylinder pressure. Published data at 5000 RPM has been used to determine possible input values to influence the SPV580 model. Cylinder pressure diagrams recorded during development of the

The combustion process is modelled by the application ofa Wiebe [18] function. Sher [19] and Heywood and Sher [20] have further analysed the Wiebe function in relation to three spark ignition two-stroke cycle engines. The mass fraction burned at a given point in the cycle can be determined from the following expression:-

SPX500 stepped piston twin-cylinder Norton “WULF” motorcycle engine in the early 1970s by Hooper and Favill [24] were also

xðqÞ ¼ 1 exp

cð q q 0 Þ

b

Dqb

(4)

Sher found from analysis of the combustion rate of the three crankcase scavenged engines that values of b ¼ 5 and c ¼ 3 to 3.2 can be recommended for small spark ignition two-stroke cycle engines. Sher and Zeigerson [21] also performed modelling analysis on a theoretical stepped piston engine based on the prior work by Harari and Sher [22]. Values for b and c based on this work have been used in order to derive a suitable Wiebe function for the computational analysis of the SPV580 as reproduced in Fig. 3.

4. Heat release model experimentation (Wiebe model)

The work of Sher and Zeigerson [21] has been explored within this study to try to achieve a closer correlation in the absence of

Fig. 3. Stepped piston engine Wiebe function based on values derived by Sher and

Zeigerson [21].

reviewed. By interrogating recorded cylinder pressure data it has therefore been possible to influence the heat release model. Fig. 4 displays data derived from the results recorded at Norton with the work of Aziz et al. and is compared with cylinder pressure data output from the baseline CFD model without overriding the default assumptions of the heat release model. The Norton data was recorded at full load 5500 RPM.

The vertical lines in Fig. 4 indicate the port timings for the SPV580 (solid), UTM125 (dashed) and SPX500 (chain dash) engines.

Whilst the SPV580 and UTM125 engines have similar swept volumes and compression ratio, the engines may well differ significantly in other areas of their design other than port timings. The bore:stroke ratio of the UTM125 is unity whilst the SPV580 and SPX500 are over-square designs. The UTM125 is a single-cylinder engine whereas the multi-cylinder SPV580 and SPX500 use cross- over charging between twin pairs of pumping and power cylinders. All of these factors could provide possible sources of difference.

Whilst it was hoped that this exploratory work may improve the observed performance and therefore improve correlation with experimental test data, the results actually showed no improve- ment. Input of the heat release model has actually created a detri- mental effect compared with the standard inactivated heat release model (no HR) shown in Fig. 5. Reduced performance can be seen in power terms particularly above 4500 RPM and a small reduction in specific fuel consumption (SFC).

4.1. CFD model results

Fig. 6 displays model output data in terms of scavenging and charging efficiency. The general fall in charging efficiency between3000 RPM and 4500 RPM is explained further within the discussion.

As previously discussed during the experimental phases of research and development of the SPV580 it was not necessary to instrument the test engines with pressure transducers. However some analysis was performed at Norton during development of the SPX500 engine. This data is reproduced for comparative purposes in Fig. 7 with the WAVE SPV580 model. In addition to thisa 1775 cm3 V4 cylinder compression ignition stepped piston enginedesigned and built at BHE was also instrumented. Being a CI engine

Fig. 4. Comparison of cylinder pressure based on work of Aziz et al. (UTM125), Norton

(NV SPX500) and SPV580 WAVE model output without heat release model input.

Fig. 5. Influence of heat release rate on the predictions of power and specific fuel

consumption at full load.

cylinder pressure data is of no comparative use for this study, however the crossover transfer passage pressure was monitored and is shown in Fig. 7 at 3000 RPM (SPD1775).

The measured pressure data for the SPD1775 and modelled data for the SPV580 show reasonable similarities. Both engines employ reed valve controlled induction systems. The higher peak pressure for the SPD1775 corresponds to the higher design primary compression ratio selected for this engine. The SPX500 was not originally designed with reed valves and utilises piston port controlled induction. This may explain the lower minimum pres- sures prior to TDC. It is interesting to note that the SPX500 appears to achieve a higher peak pressure than the WAVE predicted pres- sure for the SPV580 despite both having similar primary compression ratio (PCR) values. The exhaust and transfer port timings are also shown in Fig. 7 for the SPV580 (solid), SPX500 (chain dash) and SPD1775 (dashed).

4.2. Comparison of experimental and theoretical performance

4.2.1. Auxiliary port fuel injection (APFI) e 95 RON gasoline

An SPV580LC engine was modified to allow installation of conventional gasoline-type electro-magnetic fuel injectors. The

Fig. 6. Modelled scavenging and charging efficiency e SPV580.

Fig. 7. Comparison of stepped piston pumping cylinder pressure SPD1775 diesel

engine, SPX500 and WAVE SPV580 engines.

objective was to achieve injection at low fuel pressure into each cylinder around bottom dead centre. This required the injectors to be located within the V of the engine. However, the presence in this area of the auxiliary crossover manifold, auxiliary transfer ports and the reed valves restricted the freedom to select an optimum location and orientation for the injectors. The ideal angle of the injector would be to target the fuel away from the exhaust port and feed directly into the combustion chamber, this is clearly impos- sible with the present design, and a compromise was therefore made. These constraints and the final design are illustrated in Fig. 8.

Fig. 8. Injector location e SPV580LC (image courtesy of Bernard Hooper Engineering Ltd).

To establish a datum against which to assess the performance using injected kerosene JET A-1, data with carburetted or injected95 RON gasoline was secured [5,9,10e14]. As can be seen from Fig. 9APFI SFC is inferior to the results obtained with a carburettor. It should be noted that when employing APFI, the need to provide a rapid rate of fuel atomisation is essential for optimum fuel effi- ciency. A carburetted fuel supply has a much longer time period available for fuel/air mixing and delivery, since the fuel supply is well upstream of the combustion chamber.

Countering this disadvantage with port fuel injection however, is the ability of the engine to provide stratified charging. This is shown graphically in Fig. 9. With a carburettor, fuel and air mixture is supplied to the main and auxiliary transfer ports. With applica- tion of port injection, air only is supplied to all the transfer ports, with the fuel being injected into the air flow from the auxiliary transfer port. The auxiliary port air and fuel is steeply inclined relative to the cylinder bore towards the combustion chamber. Air only flowing from the main transfer ports provides a “wall” between the incoming auxiliary port fuel supply and the out flowing exhaust gas from the previous cycle, thus only air is lost to the exhaust port, as opposed to fuel and air in non-stratified engines.

It should however be remembered, that with the current limi- tations on injector installation, if high fuel droplet velocity is present from the injector, it is possible that the fuel could penetrate the air “wall” and exit through the exhaust port. Furthermore if fuel atomisation is poor then droplet mass will be high and the momentum of the droplets will prevent satisfactory deflection of the fuel charge towards the combustion chamber.

A range of SPV580 models were developed with a pulse width injector located in each auxiliary port attempting to simulate the experimental work with injectors located as shown in Fig. 9. Initial models used indolene fuel with subsequent analysis using a fuel file created to replicate the thermodynamic properties of kerosene JET A-1. CFD model output data using APFI of indolene is shown in Fig. 9. For comparative purposes the experimental data recorded with carburetted gasoline and APFI is also shown.

In terms of power relatively good correlation can be seen in Fig. 9 with APFI at 3000, 4000 and 4500 RPM, however this does not continue at the design maximum power speed confirmed by experimental testing. Maximum power with APFI was recorded

Fig. 9. Comparison of gasoline experimental data using carburettor (CARB) and APFI

with WAVE modelled APFI of indolene e SPV580.

experimentally at 31.61 kW at 5000 RPM. The WAVE model using indolene APFI outputs a corresponding value of only 25.64 kW at the same speed. The power output is only marginally less from the model at 4500 RPM (25.11 kW). SFC correlation is disappointing with much higher values observed with the models. The minimum SFC from experimental analysis in Fig. 9 is 0.391 kg/kWh at5000 RPM.

4.2.2. Auxiliary port fuel injection (APFI) e kerosene JET A-1Alternative fuels can be created within WAVE via input of fuel

thermodynamic properties including lower heating value, density, composition, entropy of formation, specific heat and heat of vaporisation. Kerosene JET A-1 data was collated from a number of sources notably from discussions with Shell and data published by Goodger [25,26], CRC [27], Taylor [28], Annamalai et al. [29] and Lapuerta et al. [30].

In 1975 Goodger [25] stated that kerosene fuels display average properties similar to dodecane (C12H26) and tridecane

(C13H28). In1980 Goodger [26] further stated that on average kerosene equatesapproximately to C13H25.5. These observations have therefore

been applied to the composition values within the simulations. Indolene has been used as a comparable fuel to gasoline.

4.2.3. Effects of compression ratioThe SPV580 gasoline engine has a trapped compression ratio

(TCR) of 6.75:1. Initially the ratio was reduced via shimming of the cylinder-crankcase joint face to give 6.0:1. Susceptibility to knock was still apparent with full load operation at 6.0:1. It was therefore decided to further reduce the ratio to 5.7:1.

Simulation data from APFI of JET A-1 at TCR of 5.7:1 and 6.0:1 is compared in Fig. 10 with experimental test data recorded with the SPV580 operating at 5.7:1 TCR and a JET A-1 fuel pressure of 5.5 bar.

From Fig. 10, the modelled output shows a slightly higher power at all speeds with 6.0:1 TCR. Maximum power modelled at 6.0:1TCR is 23.59 kW at 5000 RPM corresponding with 23.46 kW at 5.7:1. A reasonable correlation with test data can be seen however delineation occurs at higher model speeds and SFC correlation is relatively poor except at 3000 RPM.

4.2.4. Direct injection (DI) e 95 RON gasolineThe results of full load simulations at varying trapped air fuel

ratios of 14.7:1 and a rich operation at 13.3:1 is shown in Fig. 11 with test data recorded using carburettor fuelling.

Fig. 10. Effect of varying TCR comparing experimental data (5.5 bar JET A-1 5.7:1 TCR)

and WAVE modelled APFI of JET A-1 with 6.0:1 and 5.7:1 TCR e SPV580.

Fig. 11. Comparison of gasoline experimental data with WAVE modelled DI of indolene

e SPV580.

At 6000 RPM with a relatively rich AFR of 13.3:1 a peak maximum power of 28.38 kW can be seen. At stoichiometry the maximum power level is 27.10 kW. SFC is lowest at 14.7:1 AFR at all speeds with a minimum of 0.283 kg/kWh at 3500 RPM rising to0.312 kg/kWh at 6000 RPM.

4.2.5. Direct injection (DI) e kerosene JET A-1Models were developed to explore the DI performance using

JET A-1 at the two key experimental TCR levels. The summarised results are presented in Fig. 12.

The highest power observed from the JET A-1 DI models of26.61 kW at 6000 RPM was with 6.0:1 TCR. At 5.7:1 this reduces slightly to 26.53 kW. Power data at other speeds in Fig. 12 is also reduced with the lower TCR. The SFC using JET A-1 is also lower at6.0:1 TCR at all speeds with a minimum at 3500 RPM (0.297 kg/kWh). Using 5.7:1 TCR the minimum full load SFC increases to0.302 kg/kWh at 3500 RPM.

4.2.6. Combustion knock simulationIn order to detect occurrence of knock models developed by

Douaud and Eyzat [31] have been used based upon Fuel RON values.

Fig. 12. Effect of TCR using WAVE modelled DI of JET A-1 -SPV580.

Contact with Shell was made to see if values were available for JET A-1. Unfortunately no representative values have yet been derived as it is not relevant for the normal use of JET A-1 in turbine engines and operation in SI engines is in its relative infancy. However using correlations derived by Kalghatgi [32] a potential RON value of 58.6 has been calculated based on a Cetane value of 30. This value was therefore used in the sub-model to attempt to identify knock occurrence. At high operating speeds no knock was observed, however an example occurring at 1000 RPM is shown in Fig. 13.

In Fig. 13 the rapid cylinder pressure rise can be seen

occurring from approximately 7 ATDC. The model is operating at full load with high 6.5:1 TCR. For the intended application this load condi- tion would not in fact be possible however the evidence of knock prediction is demonstrated.

5. Discussion

The experimental test phases [5,10,12] and [13] demonstrated good results exploring kerosene fuelling of spark ignition engines. This work demonstrated the feasibility of the operation of UAV engines on lower volatility fuels. It has been observed that at low to medium speed operation an increase in power can be achieved using kerosene JET A-1 when compared with baseline gasoline test data. At maximum power speeds however the general observation has shown a small power loss.

The SPV580 engine has demonstrated power levels using kerosene JET A-1 within 5e10% of recorded gasoline levels. The further advantage of being able to operate on fuel with no added lubricant offers further advantages over many crankcase scavenged engines that require addition of a separate metering pump for heavy fuel operation.

Cruise or part load SFC is of prime importance for UAV endur- ance. Relatively high SFC can be acceptable at full load operating conditions as this is normally only required for vehicle launch and rapid climb manoeuvres. Prior part load testing of the SPV580 has demonstrated good SFC levels [9,10] and [13] with increased SFC recorded as low as 5.3% using the SPV580. Whilst this is still an increase in required mission fuel payload it nevertheless offers feasible operation and achievement of military single fuel policy objectives.

The application of computational simulation of the engines studied using Ricardo WAVE has demonstrated some interesting results. Simulation of crankcase scavenged engines has been per- formed, however the replication of a pump charging system to

Fig. 13. Cylinder pressure showing combustion knock evidence using JET A-1 (full load

1000 RPM).

provide a representative stepped piston engine model proved considerably more challenging as discussed earlier in this paper. These models however have been developed to allow simulation of the subject engine to a reasonable level of correlation within the confines of assumptions made in the absence of certain test parameters. These parameters include absence of cylinder pressure data and the ability to accurately define combustion profiles, piston surface and combustion chamber temperatures, heat transfer and scavenging details. However, whilst some reasonable correlation of power output has been observed, the same cannot be reported in terms of SFC. Apart from cases where direct injection has been modelled, the SFC levels modelled with auxiliary port injection has repeatedly shown considerably higher levels than those measured during experimental tests. It is not possible to accurately model the scavenging process during the port open period around BDC. This is clearly the period where charge short circuiting is likely to occur when both the transfer and exhaust ports are open simultaneously. Furthermore it would appear that it is not possible to replicate stratified charge operation, successfully applied to stepped piston experimental test engines, within the architecture of WAVE. If one considers for example the case of auxiliary port injection, the injector position as detailed in Fig. 9 is naturally a critically important factor with this configuration due to the proximity of the exhaust port opposite the injector. Whilst it is possible to locate the injector close to the end of the auxiliary port duct within the model, it is not possible to define the directionality of the flow from the injector. This could explain the poor SFC observed from the models attempting to simulate APFI. It is also not possible to accurately define the entry aspect of transfer ports and the desire to focus incoming charge away from the exhaust port. This may therefore further explain high levels of SFC simulated with APFI models.

Fig. 14 shows the WAVE modelled pressure time history for the SPV580 displaying cylinder, inlet, pump cylinder, transfer port and exhaust port pressures. The pumping cylinder pressure corre- sponds with the neighbouring pumping cylinder that is

utilised to charge the power cylinder. This explains the 180 phase difference between the pump pressure and the transfer port pressure. During the open cycle period when the exhaust and transfer ports are open it is possible to boost the charging efficiency of the cylinder design by effective manipulation of the forward and reverse pressure waves occurring in the exhaust system. By effective system design it is possible to draw incoming fresh charge into the exhaust port as well as into the cylinder where the charge is required. If the engine designer can create an exhaust system that provides suitably timed reverse pressure waves to arrive back at the exhaust port just prior to exhaust port closure then it is possible to force the “short-circuited” charge back into the cylinder where it can be used to positive effect by enhancing cylinder charging. This method has been successfully applied to looped scavenged two-stroke engines as a form of “ram” charging of the cylinder. The circled area in Fig. 14(i) during the period during the phase of transfer port and exhaust port closure demonstrates this critical phase. It can also be seen at 3000 RPM in Fig. 14(i) that during the circled critical phase of port overlap that a strong positive exhaust pressure pulse is present. This demonstrates a desirable condition where the exhaust pulse will provide resis- tance to the further flow of fresh charge into the exhaust and will act to return any short circuited fresh charge back into the cylinder. This may explain the good results modelled at 3000 RPM in terms of charging efficiency and power output.

Fig. 14(ii) exhibits a similar analysis but this time at 4000 RPM. It can be seen that in this case a reverse effect is occurring. Here there is a strong negative exhaust pulse. This will serve to draw further charge from the cylinder and the closing transfer ports

into the exhaust thereby ejecting this part of the new charge directly into

Fig. 14. WAVE predicted pressure time history comparison at (i) 3000 RPM and (ii)

4000 RPM.

the exhaust system. This may explain the severe fall in charging efficiency observed in Fig. 6 and the general fall in model power output from 3000 RPM to 4000 RPM. It should be pointed out that other stepped piston engines do not exhibit this drop in power as demonstrated by the data published by Hooper et al. [14].

For direct injection models, losses due to short circuiting of incoming charge, were modelled as lost air charge and much smaller quantities of fuel. The ability to model direct injection within WAVE has enabled an analysis of potential SFC with this fuel delivery configuration. The levels predicted appear to correspond with values that may be expected from a DI two-stroke engine and support arguments to explore this type of fuel system further in order to maximise thermal efficiency.

6. Conclusions

An initial exploration of the computational simulation of the SPV580 stepped piston engine has been performed in order to further investigate the results recorded from dynamometer testing.

Some good correlation has been observed however additional work is required to improve correlation further. The following conclu- sions can be drawn from the work to date.

(1) From prior experimental test work power output levels within5e10% of those recorded with 95 RON gasoline have been observed with a maximum power recorded of 30.47 kW at5250 RPM using kerosene JET A-1.

(2) The potential results of direct injection have been simulated providing output data that appears to correspond with results that could be expected from the SPV580 engine.

(3) The minimum predicted full load SFC has been observed with direct injection. Minimum full load levels using indolene fuel are 0.283 g/kWh at 4500 RPM and 0.273 g/kWh at 3000 RPM. This equates to a thermal efficiency of 29.58% and 30.67% respectively.

(4) Using JET A-1 the minimum observed full load SFC was also observed using direct injection with values of 0.315 kg/kWh at4500 RPM and 0.310 kg/kWh at 3000 RPM. In terms of full load thermal efficiency this equates to 26.58% and 27.01% respectively.

Acknowledgements

The author would like to acknowledge the support provided by UK Ministry of Defence during the design and experimental phases of the work reported and Ricardo for the provision of WAVE for exploration of theoretical simulations. The advice and information provided by Shell UK in support of the study is also gratefully acknowledged.

Appendix

NotationAFR air:fuel ratiob constant (form factor)c constant (efficiency parameter)Cf friction coefficientCm elative heat transfer area scaling factorcp gas constant pressure specific heat (J/kg K)D cylinder diameter (m)FMEP friction mean effective pressure (bar)h heat transfer coefficient (W/m2 K)P instantaneous gas pressure (bar) Pmax maximum cylinder pressure (bar)

Pr Prandtl numberRPM engine speed (r/min)S engine stroke (m)T instantaneous gas temperature (K)U gas velocity (m/s)vch characteristic gas velocity (m/s)x(q) mass fraction burned at crank angle q ( )

Dqb duration of combustion ( )q crank angle ( )q0 crank angle at start of combustion ( )r gas density (kg/m3)

AbbreviationsAPFI Auxiliary Port Fuel Injection AVGAS AViation GASoline fuel AVTUR AViation TURbine fuelBDC Bottom Dead CentreCFD Computational Fluid DynamicsCI Compression Ignition

DI Direct InjectionEC Exhaust port Closure EO Exhaust port Opening HR Heat ReleaseJET A-1 Commercial aviation grade kerosene fuelJP5 Jet Propulsion 5 fuelJP8 Jet Propulsion 8 fuelNATO North Atlantic Treaty OrganizationPCR Primary Compression Ratio RON Research Octane Number SFC Specific Fuel ConsumptionSPD1775 Stepped piston diesel 4 cylinder 1775 cm3 engineSI Spark IgnitionSPX Stepped piston crossover systemSPX500 Stepped piston twin-cylinder 497 cm3 engine developed

at Norton/BHETC Transfer port ClosureTCR Trapped Compression RatioTDC Top Dead CentreTO Transfer port OpeningUAV Unmanned Air/Aerial Vehicle

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