Hybrid_pneumatic_ICE_2011

Type Paper Number Here One Dimensional Modeling and Experimental Validation of SingleCylinder Pneumatic Combustion Hybrid Engine Pascal BREJAUD, Pascal HIGELIN, Alain CHARLET, Yann CHAMAILLARD (Do NOT enter this information. It will be pulled from participant tab in MyTechZone)Affiliation (Do NOT enter this information. It will be pulled from participant tab in MyTechZone) Copyright © 2011 SAE InternationalABSTRACTThe objective of this paper is to present and to validate a numerical model of a single-cylinder pneumatic-combustion hybrid engine. The model presented in this paper contains 0-D sub-models for non-spatially distributed components: Engine cylinder, Air tank, wall heat losses. 1-D sub-models for spatially distributed components are applied on the compressive gas flows in pipes (intake, exhaust and charging). Each pipe is discretized, using the Two-Steps Lax-Wendroff scheme (LW2) including Davis T.V.D. The boundaries conditions used at pipe ends are Method Of Characteristics (MOC) based. In the specific case of a valve, an original intermediate volume MOC based boundary condition is used. The numerical results provided by the engine model are compared with the experimental data obtained from a single cylinder prototype hybrid engine on a test bench operating in 4-stroke pneumatic pump and 4 stroke pneumatic motor modes. In each mode, the prediction of the mass flow rates, amplitude and timing of the charging pipe waves are satisfactory, without using any discharge coefficient. Indicated work and p-V diagrams are similar between simulation and measurements in the case of pneumatic pump mode. For the pneumatic motor mode the model underestimates cylinder pressure during the charging process.INTRODUCTIONFigure 1: Schematic representation of a Single CylinderThe concept of the pneumatic hybrid engine (1) (2) is based on a conventional internal combustion engine in which one additional valve, called the charging valve, connects the combustion chamber to an air tank. See Figure 1. In order to reduce costs, theintake and exhaust valve remain camshaft-driven and only the charging valve is moved by a fully variable actuator (3). As a result, thenew major modes allowed by this additional valve (pneumatic pump, pneumatic motor and supercharged) remains 4-Strokes cycles:Page 1 of 16Page 2 of 16 ∙Pneumatic Pump : This mode transforms vehicle kinetic energy into a re-usable pneumatic energy. The cycle begins with a conventional intake stroke. The charging valve opens during compression strokes as soon as the in-cylinder pressure reaches that of the charging port. The valve closes shortly after T.D.C, and the cycle finishes with conventional expansion and exhaust strokes.∙Pneumatic Motor : This mode is used to transform pneumatic energy into mechanical work without any combustion. The cycle begins with classical intake and compression strokes. The charging valve opens when approaching TDC, and remains open during part of the expansion stroke in order to maintain pressure at a high level and generate the desired indicated work. The cycle finishes with a conventional exhaust stroke.∙Pneumatic Supercharged mode : This mode is used to compensate for the "turbo lag" effect, while temporarily boost the engine during a strong acceleration transient. The charging valve opens shortly after intake valve closure, and closes when the desired additional air mass has flowed into the in-cylinder. The end of the cycle remains the same as a conventional four stroke one.Several previous publications (4) (5) (6) have shown that the concept of a pneumatic-combustion hybrid engine could provide benefits in terms of fuel consumption and greenhouse gas emissions that are comparable with those of an electric hybrid vehicle (7) but with lower weight and presumably lower cost (6). To improve the performance and efficiency of the hybrid pneumatic concept, the development mainly consists in determining the timing of opening and closing of the charging valve, for a given operating mode (pneumatic motor mode, pneumatic pump mode, supercharged mode) and for a given objective (maximization or nominal control of mechanical work, maximization or nominal control of mass or enthalpy transferred, maximization of global efficiency etc.). It has been experimentally shown (8) that large amplitude pressure waves (near 2 bars of amplitude) are created inside the charging pipe as a result of the charging valve opening and closure. In order to take this pressure wave phenomenon into account a 1D simulation platform is needed.However, the way in which gas transfer between the cylinder and the Air tank is performed depends on many parameters, such as: charging pipe length and cross section, air tank mean pressure, air tank temperature, charging valve hydraulic diameter and number, charging pipe roughness, charging valve lift. The large number of parameters to be explored for each operating mode and each objective makes optimization almost impossible to perform experimentally. A reliable, flexible, and fast numerical 1D model is therefore needed in order to develop control laws. Two previous studies (9) (10) have used a 1-D commercial simulation software for engine simulations, but do not present any comparison between experimental and simulated results for wave timings and amplitude, mass flow rates etc. From our point of view, the use of such commercial code for a pneumatic hybrid engine poses one problem and leaves one question unanswered. The problem is that the software requires a precise knowledge of the discharge coefficient versus lift of each simulated valve (11). These coefficient arrays have to be determined by experiments thus a prototype cylinder head is required. An estimated discharge coefficient will therefore be more or less erroneous, leading to serious discrepancies between simulations and future measurements. The unanswered question is that the valve boundary condition implemented in this kind of software is seldom validated on pressure ratios as high as those encountered in a pneumatic hybrid engine (up to 22). To circumvent the previous points, authors have developed their own MOC valve boundary condition that obviates in many cases the use of a Coefficient of discharge and has been experimentally validated on a valve gas dynamics test bench (12). The objective of the paper is now to present the complete engine model that embeds this valve boundary condition, and to compare experiments versus simulations.The paper is organized as follows: Firstly, 0D components models are described: the cylinder and the air tank model. Secondly, the 1D sub models for spatially distributed components are described: the LW2 numerical scheme and the simplified TVD scheme proposed by Davis are presented, followed by the Intermediate Volume Valve boundary condition. Thirdly the results produced by the complete single cylinder hybrid engine are presented and compared to measurements obtained with a prototype engine.QUASI-DIMENSIONAL COMPONENTS OF THE ENGINE MODEL AIR TANK:The air tank thermodynamic state shown on Figure 1is determined using the first law of thermodynamics:.i dU W Q h dm δδ=++(1)Page 3 of 16 Because the tank has a constant volume, 0WpdV δ=-=. According to Joule's law v v dU mc dT c Tdm =+, with Tthe temperature of the gases contained in the Tank. The air tank is assumed to be adiabatic, then 0Q δ=. For a perfect gas, the specific enthalpy is given by i p i h c T =, with i T the temperature of the flowing gases. Finally, a re-arrangement of Eq. (1) gives:().p i v vc T c T dmdT mc -=(2)The air flow mass dm is determined at the open end boundary (see corresponding section of this paper) using the method of characteristics (M.O.C). A classical numerical integration then provides the total air mass inside the tank. Using Eq. (2), the elementary temperature rise in the tank is determined. Applying a numerical integration, the new tank temperature is computed. Finally, using the perfect gas law (pV=mrT), the mean tank pressure is determined.ENGINE CYLINDERThe cylinder is the central part of the system. It exchanges heat and mass with the outside. Its thermodynamic state is determined using the first law of thermodynamics (Eq (1)).W is the mechanical work generated by the moving walls, Q is the total heat, m the mass and H is the enthalpy. Subscript i indicates an intake flow, c indicates a charging pipe flow and e indicates an exhaust flow. Then:i c e dm dm dm dm =++(3)Each mass flow rate is determined using a valve boundary condition (See corresponding section of this paper). A classical numerical integration technique then yields the new total mass inside the cylinder.The total heat Q is the sum of a combustion heat c Q and convective heat transfer through cylinders walls w Q .c w Q Q Q δδδ=+(4)w Q δ and c Q δ are respectively determined using the Woshni (13) and Wiebe (14) models. These classical correlations are notdetailed in this paper. Re-arranging Eq (1) finally gives:()p k k v vpdV Q c T dm c TdmdT mc δ-++-=∑(5)A numerical integration gives the new temperature of the cylinder. The volume is determined from the engine geometry and crankshaft position (15). The pressure is then calculated using ideal gas law.Page 4 of 16 ONE-DIMENSIONAL COMPONENTS OF THE ENGINE MODEL FUNDAMENTALSFigure 2: Control VolumeFlows, in each pipe, are assumed to be mono dimensional. The ideal compressive gas is assumed to be inviscid. Three conservation laws are written for a control volume, respectively the mass, momentum and energy. Figure 2 shows the flow of an ideal compressible gas through an infinitesimal section of pipe. ,,,,p u t x ρand F are respectively the pressure, the density, the velocity , the time, the space and the cross section area .Continuity equation: The rate of change of mass within the control volume is equal to net mass flow rate through the element. This principle leads to (16) (17):0u u dF t x F dxρρρ∂∂++=∂∂(6) Momentum equation: The rate of change of momentum within the control volume is equal to the sum of the forces applied on thecontrol volume. In order to model the friction between the flow and the wall, a shear stress w τ, opposed to the flow, is introduced. Expressing the shear stress in terms of the pipe wall frictionf gives 212w u f τρ=. D is the diameter of the pipe. 220u u p u dFG t x F dxρρρρ∂∂++++=∂∂(7) 14||2G u u f D=(8)Energy equation: The energy equation is derived from the first law of thermodynamics applied to the control volume. Heat transferper unit mass of gas q has been introduced to take into account the heat transfer between the pipe and the fluid. q is determined from an approximate treatment (16), where w T and g T are the temperature of the pipe wall and gas respectively. 0e and 0h arerespectively the stagnation specific energy and enthalpy.0000e uh uh dFq t x F dxρρρρ∂∂++-=∂∂(9) 2()1w g fu Rq T T D γγ=--(10)Page 5 of 16 TWO-STEP LAX-WENDROFF SCHEME + T.V.DFigure 3: 2 steps Lax Wendroff (Richtmeyer) stencilThe three previous governing equations can be favorably rewritten in vectorial form:()()0W F W C W t x∂∂++=∂∂(11) 0[,,]T W u e ρρρ= is the state vector of conservative variables: mass, momentum and total energy20()[,,]T F W u u p uh ρρρ=+ is the flux vector. Vector ()C W is the correction vector. It takes into account section variation,pipe wall friction and wall heat transfer:200(ln )()u d F C W u p G dx uh q ρρρρρ⎡⎤⎡⎤⎢⎥⎢⎥=++⎢⎥⎢⎥⎢⎥⎢⎥-⎣⎦⎣⎦(12) Eq (11) can be solved by a two-step Lax-Wendroff method (18), as shown in Figure 3. First step :12111121()())224 n n n n n n ni i i i i i i t t WW W F W F W C C x +++++∆∆⎡⎤⎡⎤⎡⎤=+---+⎣⎦⎣⎦⎣⎦∆(13) 12111121()()) 224n n n n n n ni i i i i i i t t WW W F W F W C C x +----∆∆⎡⎤⎡⎤⎡⎤=+---+⎣⎦⎣⎦⎣⎦∆(14)Second step :111221122() ())n n n nn ii i i i t W W F W F W tC x ++++-⎡⎤∆=---∆⎢⎥∆⎣⎦(15)Page 6 of 16 Because the LW2 scheme is not Total Variation Diminishing (T.V.D) (19) spurious oscillations appear at shocks anddiscontinuities. A flux limiter has to be added to the original LW2 scheme. The simplified T.V.D scheme of Davis (20) has been used. The advantages of this scheme are that it is symmetrical (i.e. does not depend on the direction of travel of the characteristics), and that it is easily implemented in the calculation process:111221111111111222222()())()()()()()()n n n nn n n n nii i i i i i i i i i i i i i i i t W W F W F W tC K r K r W W K r K r W W x +++++--++--++--+-++--⎡⎤⎡⎤⎡⎤∆=---∆++-++-⎢⎥⎢⎥⎢⎥∆⎣⎦⎣⎦⎣⎦(16)The functions i K +-and ir +-are not detailed is paper, and the reader is referred to the original paper (20).OPEN END BOUNDARY CONDITIONFigure 4: Open End Boundary ConditionThe open end boundary condition used at the pipe ends is Method Of Characteristic based and uses the C.I.R Method (21). Figure 4 depicts the general context of this boundary. Non dimensional notations from Benson are used (16; 17). Subscripts ‘ref ’ and ‘ext ’ are respectively the reference and the exterior state.Outflow from pipe to cylinder:In this case, the characteristic line C +and path-line 0C exist as depicted in Figure 4.Three primitive quantities at the boundary are unknown :1n iρ+,1n iu + and1n i p +.First, using the CIR method, the non-dimensional Riemann invariant along the C +characteristic line is determined111122n nn n in L L i i refA U A U a αγγλ++--==+=+(17)If the flow is subsonic, the pressure at the boundary is assumed to be equal to the exterior pressure1n i ext p p +=(18)If the flow is subsonic, then contraction downstream the throat is assumed to be isentropic:2111n n i iext ext pa p a γγ++-⎡⎤=⎢⎥⎣⎦(19)Re-arrangement of Eq.(18) and Eq.(19) gives :Page 7 of 16 1n i ext A A +=(20)Using equations (17)(18)(20) the subsonic state at the boundary condition can be solved. The obtained solution has to be checked as a subsonic one. In the contrary case, the constant pressure and homentropic contraction hypotheses are no longer valid. The flow is now assumed to be sonic at the throat:11n n i i U A ++=(21)The homentropic flow along the 0C path line gives :()()11n ni s n n isp p γγρρ++=(22)Using equations (17)(21)(22) sonic state at the boundary condition can be solved.Inflow from cylinder to pipe:In this case, the pathline 0C no longer exists. Three primitive quantities at the boundary are unknown: 1n i ρ+,1n iu + and1n i p +.If the flow is subsonic, the characteristic line C +exists, then :111122n nn n in L L i i refA U A U a αγγλ++-==+=+(23)Gas contraction upstream the throat is assumed to be homentropic:221111()n n i ii ext ext pa A p a γγγ++--⎡⎤==⎢⎥⎣⎦(24)The steady state total energy conservation equation gives :12121()()2n n ext i i A A U γ++-=+(25)Using equations (23)(24)(25) the inflow subsonic state at the boundary condition can be solved. The solution obtained has to be checked as a subsonic one. In the contrary case, the characteristic line C + is no longer valid, and Eq (23) cannot be used and must be replaced by the sonic flow condition at throat, given by Eq (21). The sonic inflow state is then computed using equations (21)(24)(25).Page 8 of 16 Entropy correction for gas inflowing into pipe:Figure 5: (a) Benson Entropy Level (b) Entropy correction at BoundaryIn the case of gases inflowing into the pipe at the boundary condition, the M.O.C. resolution used in the previous section assumes an homentropic flow. This assumption means, firstly, that gases located in the first node of pipe are at the same entropy level as those in the cylinder, and secondly, that there is no entropy creation across the valve throat. Needless to say, these ideal conditions are never encountered in a real flow, thus making entropy correction necessary. The change of Riemann variable along the characteristic line C - due to a change of entropy level is defined by the following equation, see Benson (16):,,,,2in c out c Ac An Ain c in n A Ac A A dA d AA A λλλλλ+⎛⎫⎛⎫-=-=⎪ ⎪⎝⎭⎝⎭(26)As the three quantities ,in c λ, ,out c λ and Ac A refer to a new time step level 1n +, Eq (26) has to be solved by an iterativemethod. In chapter 7 of his book, Benson (16) proposes an algorithm to converge previous equation, based on the estimated entropy level Ac A .INTERMEDIATE VOLUME VALVE BOUNDARY CONDITIONFigure 6: Intermediate Volume Valve Boundary ConditionThe original MOC valve boundary condition (see Figure 6) has been defined and experimentally validated in a previous work from authors (12) . The main interests of this boundary condition are first, to avoid any convergence problem or non-physical shock that may be encountered while using of the classical constant pressure model from Benson, and second, to obviate the use of a Coefficient of discharge. See (12). The boundary is constituted by an intermediate volume that connects the pipe and the cylinder. This volume exchanges mass through two ports. The first one is a MOC Open End Boundary Condition (see previous section) thatconnects the intermediate volume and the pipe. The second one is the valve port that connects the intermediate volume to the cylinder.Page 9 of 16 Mass flow rate through the valve port:The mass flow rate through the valve port is calculated with the classical Barre de Saint Venant equation, governed by the pressure ratio at the throat. Four cases must be distinguished. The flow can be in the forward or the reverse direction (back flow) and it can be sub or supersonic. Eq (29) presents the critical pressure ratio (c r ) at the throat which determines the type of flow. If thepressure ratio at the throat, defined as dpup r p =, is lower than the critical ratio, the flow is supersonic, otherwise the flow is subsonic. Because the pressure ratio p r must be lower than unity, u p and d p are defined as:1( ,)n u iv ext p max p p +=(27) 1( ,)n d iv ext p min p p +=(28)121c r γγγ-⎛⎫= ⎪+⎝⎭(29)Eq (30) and Eq(31)give the mass flow rate at the throat respectively for a subsonic and a sonic flow. v Cd is a possible dischargecoefficient at the valve if necessary:1vd v t u dm p Cd F dtp γ⎛⎫=⎪⎭12(1)0.521v v t dm Cd F dt γγγ+-⎛⎫= ⎪+⎝⎭(31)The sign ofvdm dtmust be adapted to the direction of the flow: the sign is set positive if the flow enters the intermediate volume. Mass flow rate through pipe port:The mass flow rate through the pipe port of the intermediate volume (p subscript) is determined using the Open End Boundary Condition presented before. This mass flow rate is determined using Eq (32). It can be seen that no discharge coefficient is used at this port:11**p n n i i i dm F u dtρ++=(32)Iterative resolution scheme:Because each mass flow ratep dm dtandv dm dtdepends on the pressure level of the intermediate volume at the new step time (1n iv p +), an iterative resolution is needed. Reader is pleased to refer to original paper for details (12).Page 10 of 16EXPERIMENTAL VALIDATIONTable 1 – Prototype engine dimensionsThe pneumatic hybrid engine used for experiments is a prototype single cylinder, 347 cm 3 displacement, 4-Stroke, Spark-Ignition, with hemispherical cylinder head and flat piston. Table 1 gives the general engine dimensions and Table 2 gives the specifications of the valves, pipe and actuators. In-cylinder pressure is measured by a piezoelectric pressure sensor. The intake, charging and exhaust port pressures are measured by three absolute piezo-resistive sensors. Intake air port temperature is measured by a Pt100 sensor, and exhaust and charging port temperatures are measured by two K-thermocouples. The intake and charging mass flow rates are measured with two Mass flow controllers. The exhaust pressure is regulated by a control valve. The intake and charging air can be warmed up by two resistive heaters. Two plenums of respectively 6 and 30 liters volume are mounted on intake and exhaust lines. Temperature controls are implemented on both water and oil circuits. Two operating modes of the engine are studied: 4-stroke Pneumatic Pump and 4- stroke Pneumatic Motor.Table2 –Valves and pipes specificationsEach simulation performed included 4 complete engine cycles, starting from null speed steady state inside each pipe. All the numerical results presented were taken from the third simulated cycle, to ensure that non-stationary states were sufficiently present in each pipe. The discharge coefficient of all the valves was set to unity. Woshni's heat loss coefficient 1C (13) was fitted on pure compression measurements. It should be noted that a strong hypothesis has been made by supposing a constant wall temperature along a given pipe (intake, exhaust and charging). Each pipe is modeled as a straight duct without any sudden enlargement and section change.Table 3 –Operating parametersFour-Stroke Pneumatic Pump mode :Table 4 –Numerical results for 4-stroke pneumatic pump modeby the entire engine model. Figures 7 and 8 respectively display the p-V diagram and the charging port pressure versus Crack Angle.Figure 7: pV diagram- Pneumatic 4-stroke pumpFigure 8: Charging port pressure versus Crank Angle- Pneumatic 4-stroke pumpThe simulated p-V diagram has a general shape close to the measured one, a point that is confirmed by the very low relative error ( -1.27 %) between the measured and simulated values of the indicated WorkiW . The simulated value of charging air mass per cycle exhibits a +4.87 % relative error compared to the measured one. This result can be considered excellent, given that the discharge coefficient was set to 1, and never have been fitted.Page 11 of 16Page 12 of 16 Analysis of the charging port pressure vs. crank Angle (see Figure 8) shows that the simulated pressure wave is very close to the measured one in terms of amplitude and timing. This is a key point with regard to our objective, which is to be able to predict the optimum timings of opening and closing of the charging valve.Figure 9 : Intake and exhaust charging port pressure versus Crank Angle- Pneumatic 4-stroke pumpThe simulated intake air mass presents a deviation and a relative error respectively of +61 mg/cycle and 19.3 %, in respect to measurement. This poor result is due an inappropriate size and form of the intake plenum. Thus the waves existing inside the upstream and downstream pipes of the plenum were not sufficiently decoupled. This affirmation is clearly highlighted by Figure 9. The simplest way to solve this problem would have been to improve the plenum, in order to sufficiently decouple the upstream and downstream waves.Before regarding to the exhaust air mass per cycle, it is important to note that to obtain a correct simulated value, it is necessary to have accurate values for intake and charging air mass per cycle. All errors are echoed on exhaust air mass value, because 0intake charging exhaust m m m ++=. When this fact is taken into account, the observed deviation of -50 mg/cycle between simulated and measured exhaust air mass is logical, because of the deviations observed on intake and charging air mass respectively of +61 mg/cycle and -11 mg/cycle. Moreover, an analysis of the exhaust port pressure versus Crank Angle (Figure 9) shows that the amplitude and decay of the pressure wave is correctly described by the model, but that a phase shift is gradually installing. Authors assume that this phenomenon is due to a non-negligible convective heat transfer existing along the charging pipe which has not been taken into account by the model.Because the main goal was to correctly modelize the charging pipe behavior, authors have not spent more time to solve those issues.Four-Stroke Pneumatic Motor mode:Table 5 –Numerical results for 4-stroke pneumatic motor modePage 13 of 16Figure 10: p-V diagram- Pneumatic 4-stroke motorFigure 11: Charging port pressure versus Crank Angle- Pneumatic 4-stroke motorTable 5 presents the numerical results for a representative measured cycle on the test bench, and numerical results produced by the entire engine model. Figures 10 to 12 respectively display the p-V diagram, the charging, the intake and the exhaust port pressure versus Crack Angle.Figure 12: Intake and exhaust port pressure versus Crank Angle- Pneumatic 4 stroke motorPage 14 of 16 First, an analysis of the charging port pressure vs. Crank Angle (Figure 11) shows that the simulated pressure wave is very close to the measured one in terms of amplitude and timing. This is a key point for a future valve actuation timings optimization. Moreover, the analysis of the charging air mass per cycle shows a relative error of only -3.89% between air simulated and measured is observed (Table 5). Analysis of Figure 12 leads to similar conclusions as for the pneumatic pump mode. The intake plenum is inappropriate and convective heat transfer along the exhaust pipe cannot be neglected.Surprisingly, a comparison of the shapes of the measured and simulated p-V diagrams leads to less satisfactory results than for the pneumatic pump mode. The observed relative error of -17.98% between measured and simulated values of the indicated Work i W confirms this point. Visually, it is obvious that the variation occurs during the opening of the charging valve. The measured in-cylinder pressure is higher than the simulated one. Thus more energy than expected has been transferred to the in-cylinder gas. Authors conjecture that the opening of the charging valve, given of the high pressure ratio that exists between the pneumatic charging port and the cylinder, significantly changes the in-cylinder internal aerodynamic structure that produces a positive convective heat transfer process.SUMMARY/CONCLUSIONSThe present work proposes a complete 1D engine model for pneumatic-combustion hybrid engine. A comparison of experimental and simulated results with no discharge coefficient has shown good results, at the charging port, in terms of mass flow rates per cycle, pressure wave timings and amplitude for both investigated pneumatic modes. For the pneumatic 4-stoke pump mode, the predictions of Indicated Work and p-V diagram were satisfactory. But, in the special case of the pneumatic 4-stroke motor, the indicated work was under estimated by the model. The model is clearly underestimating the energy transfer during the pneumatic charging process. It has been supposed that this phenomenon is due to a convective heat transfer process created by a strong in-cylinder air motion during the pneumatic charging. Future works has to experimentally investigate the in-cylinder aerodynamic structure and the heat transfer process during the charging valve opening.As a final conclusion, the gas dynamic modeling into the charging pipe and mass flow rates through valves has been experimentally validated that allows a future charging valve timing optimization work.REFERENCES1. New cycles for automobile engines. SCHECHTER, M. 1999, SAE Technical Paper, Vols. 1999-01-0623.2. Thermodynamic cycles for a new hybrid pneumatic-combustion engine concept. HIGELIN, P. and CHARLET, A. 2001, SAE Paper, Vols. 2001-09-23.3. Modelling and optimizing two- and four-stroke hybrid pneumatic engines. DONITZ, C., et al. 2009, IMechE Part D - Journal of Automobile Engineering, Vols. Vol. 223, No. 2, pp. 255-280.4. Realizing a concept for high efficiency and excellent driveability: The downsized and supercharged hybrid pneumatic engine. DONITZ, C., et al. 2009, SAE paper, Vols. 2009-01-1326.5. Energy Management Strategies for a Pneumatic-Hybrid Engine Based on Sliding Window Pattern Recognition. IVANCO, A., et al. 2009, Oil and Gas Science and Technology, Vol. DOI: 10.2516/ogst/2009045.6. Vehicle Driving Cycle Simulation of a Pneumatic Hybrid Bus Based on Experimental Engine Measurements. TRAJKOVIC, S., TUNESTAL, P. and JOHANSON, B. 2010-01-0825 , 2010, SAE Technical paper, Vols. 2010-01-0825.7. Well-To-Wheels Use and Greenhouse Gas Emissions of Plug-in-Hybrid Electric Vehicles. ELGOWAINY, A., et al. 2009, SAE paper, Vols. 2009-01-1309.8. Investigation of Different Valve Geometries and Valve Timing Strategies and their Effect on Regenerative Efficiency for aPneumatic Hybrid with Variable Valve Actuation. TRAJKOVIC, S., TUNESTAL, P. and JOHANSSON, B. 2008, SAE Paper, Vols. 2008-01-1715.。