Int.J. Applied Thermodynamics, ISSN 1301-9724 Vol.5, (No.1), pp.1-11, March-2002

Thermodynamic Simulation of a Hybrid Pneumatic-Combustion Concept

Pascal HIGELIN*, Alain CHARLET, Yann CHAMAILLARD LME-ESEM, Université d’Orléans 8, rue Léonard de Vinci, 45072 Orléans cedex 2 E-mail: [email protected]

Abstract Although internal combustion display high overall maximum global efficiencies, this potential cannot be fully exploited in automotive applications: in real conditions, the average engine load (and thus efficiency) is quite low and the kinetic energy during a braking phase is lost. This work presents a new hybrid pneumatic- combustion engine and the associated thermodynamic cycles, which is able to store energy in the form of . This energy can be issued from a braking phase or from a combustion phase at low power. The potential energy from the air tank can then be restored to start the engine, or charge the engine at full load. The regenerative breaking and the suppression of the idling phases could provide an improvement in terms of fuel economy as high as 15% or more if combined with engine downsizing. Key words: internal combustion engine, , thermodynamic cycles, pneumatic pump

1. Introduction must be dissipated into heat through the braking system. To be able to meet future challenges for internal combustion engines in terms of pollutant One solution to this problem is to use a emissions and greenhouse effect gases emissions, hybrid powertrain combining a conventional the global efficiency over a real world driving combustion engine for easy energy storage and cycle needs to be improved. Today, internal high overall conversion efficiency, and a combustion engines display high overall reversible energy converter and storage maximum global efficiencies, topping at about subsystem. Usually the second energy converter 45% for turbocharged direct injection (DI) diesel is an and the energy storage engines. But this great potential cannot be fully subsystem is a battery. The downside of this exploited: configuration is that the electric motor, generator • The first problem is that the peak and battery are heavy and expensive. efficiencies can only be reached at full load. In In this work, we propose a new hybrid real life operation, for example city traffic, the engine concept, which combines an internal situation is even worse: full load is never combustion engine with a pneumatic energy reached, the engine is mainly idling or operating converter and storage subsystem. To lower the at very low load, and thus at very low efficiency. cost and weight and to stay with a well-known • The second problem with internal technology, the internal combustion engine itself combustion engines is that the thermodynamic has been used as a pneumatic pump or motor cycle cannot be reversed: i.e. you cannot produce when there is no combustion. air and fuel from combustion products by 2. The Concept providing reverse to the crankshaft. As a result, the kinetic energy of a vehicle cannot be The concept is based on a conventional recovered into chemical energy in negative internal combustion engine where one additional torque i.e. braking situations: the kinetic energy valve called the charging valve is connecting the combustion chamber of each cylinder to an air

* Author to whom all correspondence should be addressed.

Int.J. Applied Thermodynamics, Vol.5 (No.1) 1 tank (Schechter, 1999). The timing of the revolution: it is a two-stroke cycle. Figure 2 opening and closing of this valve is driven by an displays the idealized P-V diagram: additional . The additional valve could Cylinder charging: The charging valve is also be electrically actuated (Figure 1). opened from 1 to 3. The air tank discharges into the cylinder between 1 and 2, where the cylinder reaches the tank pressure. From 2 to 3 the High pressure air tank discharge is continued at a constant pressure (assuming the tank volume is much larger than the cylinder volume). The work of the pneumatic motor cycle can be continuously varied by adapting the timing of the charging valve closure Intake Exhaust (Figure 2, dashed lines). For low load operation, points 2 and 3 are superimposed and the corresponding pressure might be lower than the tank pressure. Expansion: From 3 to 4 the charging valve is closed and the air charge expands. The expansion stroke (2 to 4) produces the work of the cycle. To obtain a full expansion of the air charge, the timing of the charging valve closure must be optimized in order to superimpose points Figure 1: Each combustion chamber is 4 and 5. connected to a high-pressure air tank through an Exhaust: The exhaust valve opens at point additional valve and pipe 4. From 4 to 5 the cylinder discharges into the With this concept, in addition to the exhaust pipe and the potential energy of the air at conventional internal combustion engine point 4 is lost. From 5 to 1 the exhaust stroke operation, several new operating modes must be expels the remaining air into the exhaust pipe at considered (Higelin et al., 2001): atmospheric pressure. The two stroke pneumatic motor operating 2.1 Pneumatic motor operation mode leads to high torque output as well as low The hybrid pneumatic-combustion engine friction losses. During pneumatic motor can produce torque from 0 rpm, operating as a operation, no fuel is injected into the intake pneumatic motor powered by the high-pressure ports. air tank. As long as the air tank has sufficient air This operating mode can be used to pressure available, this potential energy can be completely suppress the idling operation of converted to mechanical energy on the conventional engines by shutting down the crankshaft at any time, just by opening the engine at each stop. After a stop, the engine can charging valve. be rapidly brought to normal speed and start using the high level of torque provided by the P pneumatic motor operating mode. It is well known that engine operation without idling phases can substantially improve fuel economy in typical city driving. Due to the high level of torque available during pneumatic motor operating, an alternative starting mode could be to start the engine with clutch (or automatic transmission) engaged. So, the engine operates in pneumatic mode for a few cycles and switches to combustion mode as soon as the engine speed is high enough. With this

TDC BDC strategy, the vehicle starting is brought from a V two-step process (engine starting and engine Figure 2. Idealized P-V diagram of speed going up to minimum operating speed, pneumatic motor operation for several air then transmission engaging) to a one-step charges process. In this mode, the vehicle starting process can be as smooth as with an electric The pneumatic motor cycle consists of three motor if the number of cylinders is high enough. steps: cylinder charging, expansion and exhaust. The full cycle is completed within one crankshaft 2 Int.J. Applied Thermodynamics, Vol.5 (No.1) The pneumatic motor operating mode can Compression: From 2 to 3 all the valves are also be used for low power in-city cruising with closed and the air charge is compressed up to the zero pollutant emissions. However, to get a tank pressure. If the braking torque with full air sufficient operating range, the size of the air tank mass is not high enough, the timing of the must be appropriately dimensioned. charging valve opening can be advanced to 2.2 Pneumatic pump operation obtain a cycle with a larger area (Figure 4, points 1-2-5-6-4-1). The same effect can be obtained by One of the largest weaknesses of internal retarding the opening of the charging valve, at combustion engines is that the thermodynamic the expense of a much higher maximum cylinder cycle cannot be reversed: i.e. you cannot produce pressure (Figure 4, points 1-2-7-8-4-1) air and fuel from combustion products by providing reverse torque to the crankshaft. As a P result, the kinetic energy of a vehicle using a conventional internal combustion engine cannot be recovered into chemical energy during negative torque, i.e. braking situations: the kinetic energy must then be dissipated into heat through the friction of the braking system. 8

P 3

BDC TDC V Figure 4. Idealized P-V diagram of pneumatic pump operation for three timings of the charging valve opening Air tank charging: From 3 to 4 the charging valve is open. The end of the compression stroke TDC BDC V is used to expel compressed air from the cylinder Figure 3. Idealized P-V diagram of to the air tank. The air tank charging performed pneumatic pump operation for two intake and is at nearly constant pressure, assuming the two charging valve timings volume of the air tank is much larger than the The hybrid pneumatic-combustion engine displacement of the engine. can produce negative torque, i.e. consume torque Expansion: From 4 to 1, all the valves are from the driveline during a braking phase. The closed. The compressed air that could not be kinetic energy of the vehicle is converted to pushed into the air tank expands to atmospheric potential energy in the form of pressurized air pressure when the intake valve opens (point 1). pumped into the air tank. To lower the braking torque, the closure of the The pneumatic pump cycle consists of four charging valve can be retarded to 4’ (Figure 3). steps: intake, compression, air tank charging and In this case, part of the air mass inside the air expansion. Again, the full cycle is completed tank at point 4 is swept back into the cylinder within one crankshaft revolution: it is a two- between points 4 and 4’. To enhance the braking stroke cycle. Figure 3 displays the idealized P-V torque, the expansion stroke can be completely diagram: or partially removed (Figure 4). In this case, the exhaust valve is opened at point 4 (or 4’ for a Intake: The intake valve opens at point 1. partial expansion) and closed at point 1. The From 1 to 2 the cylinder is filled with fresh air cycle then becomes 1-2-3-4-9-1 or 1-2-3-4-4’-9’- from the intake port. The intake valve is closed at 1 (see Figure 4). point 2 if the maximum air mass is needed. To reduce the air mass swept by the cycle, and The pneumatic pump operating mode can conversely the cycle braking work, the intake be used for regenerative braking. The braking valve closure can be retarded to 2’ (Figure 3). In torque as well as the mass of compressed air this case, part of the air mass inside the cylinder stored into the air tank can be continuously at point 2 is swept back into the intake manifold adjusted by varying the intake, charging and between points 2 and 2’. exhaust valve timings.

Int.J. Applied Thermodynamics, Vol.5 (No.1) 3 2.3 Supercharged engine operation (Figure 5, dashed line) comparatively to the During strong transient accelerations or atmospheric cycle (Figure 5, solid line) provides, short-term high power output, the compressed air as expected, a higher IMEP, but also a higher contents of the high-pressure tank can be used to level of maximum cylinder pressure. This higher supercharge the internal combustion engine. The pressure can be overcome by retarded supercharging pressure can be quite high and is combustion timing, at the expense of the global only limited by the compressed air tank pressure. cycle efficiency (Figure 5, dash-dotted line). Compression: The main difference between P a conventional internal combustion engine cycle and the pneumatically supercharged cycle resides in the compression stroke. This compression stroke is composed of three phases: • After the closing of the intake valve, a conventional compression phase starts, bringing the in-cylinder air to a pressure below the tank pressure (Figure 5, points 1-10). • Then, the charging valve opens, allowing additional air to flow from the high-pressure tank into the combustion chamber (Figure 5, points 10-11-12). • At the closing of the charging valve, a second compression stroke is performed (Figure 5 points 12-2’). This strategy offers three main differences TDC BDC compared to a conventional supercharged cycle: V Figure 5. Idealized P-V diagram of engine • By using the conventional intake and supercharging with air from the high-pressure exhaust stroke, the tank only has to provide an tank amount of air in excess of the cylinder charge at atmospheric pressure. Consequently, the lifetime The positive effect of this supercharging capability is that the engine design can be of the high pressure tank air charge can be optimized for the maximum power meeting 80 to greatly enhanced (2 to 4 times). 90% of the driving situations without • Unlike a conventional supercharged cycle, supercharging. The remaining 10 to 20% high where the complete air mass is in the cylinder power output operating conditions are met by when the intake valve closes, the pneumatically transient supercharging with the compressed air supercharged cycle can provide a higher high- from the high-pressure tank. Thus, the engine pressure cycle IMEP due to the dual stage can be downsized, allowing a higher mean compression stroke. Indeed, the first part of the operating efficiency over the whole range of compression stroke is performed with a lower air operating conditions. mass, i.e. lower pressure, providing additional The pneumatically supercharged cycle work to the cycle. To maximize IMEP, the consists of several steps: intake, compression 1, supercharging process should be performed as supercharging, compression 2, combustion, late as possible. The optimal strategy would be to expansion and exhaust. Figure 5 displays the close the charging valve when cylinder pressure idealized P-V diagram: reaches tank pressure, and to determine the Exhaust and intake: The pneumatic timing of the charging valve opening for the supercharged cycle is identical to a conventional desired additional air mass. four-stroke engine cycle during the exhaust • On the downside, the exhaust pressure of (Figure 5, points 5-6-7) and intake (Figure 5, the hybrid engine being higher than the intake points 8-9-1) process. pressure, the low-pressure cycle IMEP is Combustion and expansion: The negative. A conventionally supercharged engine combustion and expansion phases of the benefits from a positive low-pressure cycle. supercharged cycle (Figure 5, points 2’-3’-4’-5’) 2.4 Undercharged engine operation display the same shape as the non-supercharged cycle (Figure 5, points 2-3-4-5). The higher level The energy recovered during the braking of air mass inside the cylinder during the phases is usually not enough to provide a full air supercharged combustion and expansion phase tank charging. It is generally not possible to get

4 Int.J. Applied Thermodynamics, Vol.5 (No.1) the same state of charge (SOC) of the high- • The first compression being performed with pressure air tank at the end of a test cycle than at the full air charge, the undercharged cycle is the beginning. Therefore, to fully benefit from always performed with a wide-open throttle. In the hybrid pneumatic-combustion engine this case, points 7 and 8 as well as 1, 6 and 9 concept, an operating mode allowing the air tank (Figure 6) are superimposed, providing zero charging while in combustion mode must be pumping losses during the low-pressure cycle. added. The global IMEP is increased, as well as the global efficiency. The undercharged engine operation mode consists of taking away part of the fluid enclosed • On the downside, during the first in the combustion chamber during the compression stroke (Figure 6, points 1-10), work compression to charge the high-pressure air tank. is drawn from the crankshaft to compress the air mass that will be pumped into the high-pressure The undercharged engine cycle consists of tank. Thus, the high-pressure cycle IMEP is several steps: intake, compression 1, charging, lower than with a conventional low-load cycle. compression 2, combustion, expansion and exhaust. Figure 6 displays the idealized P-V P diagram: Exhaust and intake: The undercharged engine cycle is identical to a conventional four- stroke engine cycle during the exhaust (Figure 6, points 5-6-7) and intake (Figure 6, points 8-9-1) phase. Combustion and expansion: The combustion and expansion phases of the undercharged cycle (Figure 6, points 2’-3’-4’-5’) display the same shape as a conventional four- stroke engine cycle (Figure 6, points 2-3-4-5). The lower level of air mass inside the cylinder during the undercharged combustion and expansion phase (Figure 6, dashed line) TDC BDC comparatively to the atmospheric cycle (Figure V 6, solid line) provides, as expected, a lower Figure 6. Idealized P-V diagram of IMEP, but also a lower level of maximum undercharged engine operation cylinder pressure. To optimize the global cycle efficiency, the start of combustion can be 3. Hybrid Engine Model advanced to reach a higher peak cylinder To be able to compute the global efficiency pressure. over a test cycle, the efficiency and IMEP of Compression: The main difference between each operating mode must be expressed as a a conventional internal combustion engine cycle function of tank pressure, as well as intake, and the undercharged cycle resides in the exhaust and charging valve timing: compression stroke. This compression stroke is 3.1 Pneumatic motor composed of three phases: The transformation from 1 to 2 being at • After the closing of the intake valve, a constant volume without heat exchange (Figure conventional compression phase starts, bringing 2), application of the first law yields: the in-cylinder air to the same pressure as the tank (Figure 6, points 1-10). m2 ⋅ u2 − m1 ⋅ u1 = (m2 − m1)⋅ ht (1) • Then, the charging valve opens, allowing p air from the combustion chamber to flow to the T = t (2) 2 p − p p high-pressure tank during part of the t e + e compression stroke (Figure 6, points 10-12). γ ⋅Tt Te • At the closing of the charging valve, a V ⎛ p p ⎞ second compression stroke is performed (Figure m ⎜ t e ⎟ m2 − m1 = ⋅⎜ − ⎟ (3) 6, points 12-2’). r ⎝ T2 Tr ⎠ This strategy offers two main differences The process from 2 to 3 being isobaric at air compared to a conventional low-load operating tank pressure (Figure 2), it comes: mode:

Int.J. Applied Thermodynamics, Vol.5 (No.1) 5 m ⋅ u − m ⋅ u = W34 = −pt ⋅ (Vm − V3 ) (18) 3 3 2 2 (4) = ()m 3 − m 2 ⋅ h t − p t ⋅(V3 − Vm ) The expansion process from 4 to 1 is considered adiabatic and reversible (Figure 3): pt ⋅V3 ⎛ T2 ⎞ pt ⋅Vm m3 = −m2 ⋅⎜ −1⎟− (5) ⎜ ⎟ W41 = m1 ⋅ (u1 − u4 ) (19) r⋅Tt ⎝γ⋅Tt ⎠ cp ⋅Tt

pt ⋅ V3 The specific mass of the pneumatic pump T3 = (6) m3 ⋅ r operating mode can be written as the quotient of mass pumped into the tank (12) and mechanical W23 = −pt ⋅(V3 − Vm ) (7) work consumed (14, 16, 18, 19): m − m Since the expansion process from 3 to 4 is ζ = 2 1 (20) ⎛− p ⋅()V − V − p ⋅(V − V )⎞ considered adiabatic and reversible (Figure 2): ⎜ i 2 1 t m 3 ⎟ ⎜+ m ⋅ u − u + m ⋅ u − u ⎟ γ−1 ⎝ 1 ()1 4 2 (3 2 )⎠ ⎛ V ⎞ ⎜ 3 ⎟ T4 = T3 ⋅⎜ ⎟ (8) 3.3 Pneumatic supercharging ⎝ VM ⎠ The compression from 1 to 10 is considered W34 = m3 ⋅ (u4 − u3 ) (9) isentropic without heat exchange (Figure 5), hence: The exhaust stroke being isobaric (Figure γ−1 2), the produced mechanical work can be ⎛ V ⎞ ⎜ M ⎟ expressed: T10 = T1 ⋅⎜ ⎟ (21) ⎝ V10 ⎠ W51 = −pe ⋅(Vm − VM ) (10) γ ⎛ V ⎞ The specific work of the pneumatic motor ⎜ M ⎟ p10 = p1 ⋅⎜ ⎟ (22) operating mode can be written as the quotient of ⎝ V10 ⎠ work produced (7, 9, 10) and mass provided by the tank (3, 5): W110 = m1 ⋅ (u10 − u1) (23)

m ⋅()u − u − p ⋅(V − V )− p ⋅(V − V ) The flow through the charging valve is ξ = 3 4 3 t 3 m e m M (11) m − m modeled by a constant volume transformation 3 1 from 10 to 11 followed by a constant pressure 3.2 Pneumatic pump transformation from 11 to 12 (Figure 5): The intake stroke being isobaric and p T = t (24) isothermal at intake pressure and temperature 11 p − p p (Figure 3): t 10 + 10 γ ⋅Tt T10 p ⋅ ()V − V m − m = i 2 1 (12) 2 1 r T V ⎛ p p ⎞ ⋅ i m − m = 10 ⋅⎜ t − 10 ⎟ (25) 11 1 r ⎜ T T ⎟ 1 ⎝ 11 10 ⎠ ⎛ p ⎞ γ V = V ⋅⎜ t ⎟ (13) m ⋅ u − m ⋅u = 1 m ⎜ ⎟ 12 12 11 11 (26) ⎝ pi ⎠ = ()m12 − m11 ⋅h t − p t ⋅(V12 − V10 )

W12 = −pi ⋅ (V2 − V1) (14) pt ⋅ V12 m12 = − The compression process from 2 to 3 is r ⋅Tt considered adiabatic and reversible (Figure 3): (27) ⎛ T ⎞ p ⋅ V − m ⋅⎜ 11 −1⎟ − t 10 1−γ 11 ⎜ ⎟ ⎝ γ ⋅Tt ⎠ cp ⋅Tt ⎛ p ⎞ γ ⎜ i ⎟ T3 = Ti ⋅⎜ ⎟ (15) ⎝ pt ⎠ pt ⋅ V12 T12 = (28) m12 ⋅ r W23 = m2 ⋅ (u3 − u2 ) (16) W = −p ⋅(V − V ) (29) The tank charging process from 3 to 4 is 1112 t 12 10 considered isobaric and isothermal without heat Again, the second stage of the compression exchange (Figure 3): process is considered isentropic without heat exchange (Figure 5), it comes: T4 = T3 (17)

6 Int.J. Applied Thermodynamics, Vol.5 (No.1) γ−1 supercharged cycle in that point 10 and point 11 ⎛ V ⎞ ⎜ 12 ⎟ T2' = T12 ⋅⎜ ⎟ (30) are superimposed at tank pressure, hence: ⎝ Vm ⎠ 1 γ γ ⎛ p1 ⎞ ⎛ V ⎞ V10 = VM ⋅⎜ ⎟ (43) ⎜ 12 ⎟ ⎜ p ⎟ p 2' = p t ⋅⎜ ⎟ (31) ⎝ t ⎠ ⎝ Vm ⎠

T11 = T10 , p11 = pt , m11 = m1 (44) W122' = m12 ⋅(u2' − u12 ) (32) In this case, the undercharged hybrid The combustion process is modeled by heat pneumatic engine efficiency has the same transferred to the working gas: expression as the supercharged efficiency: Q = η ⋅ m ⋅Q (33) ⎛m ⋅(u − u )(− p ⋅ V − V ) ⎞ c f LHV ⎜ 1 10 1 t 12 10 ⎟ The first part (from 2’ to 3’) of the ⎜+ m12 ⋅()u2' − u12 + u5' − u4' ⎟ ⎜ ⎟ combustion process is modeled by a constant − p ⋅()V − V + (p − p )⋅(V − V ) η = ⎝ 3' 4' m e i M m ⎠ (45) volume process, the second part (3’ to 4’) is u mf ⋅QLHV modeled by a constant pressure transformation (Figure 5): 3.5 Conventional combustion The conventional combustion cycle has a Qv = εv ⋅Q (34) single compression stroke at constant mass (Figure 5): Q p = ε p ⋅Q = (1− ε v )⋅Q (35) γ ⎛ V ⎞ Q m ⋅ r ⋅T p = p ⋅⎜ M ⎟ (46) T = T + v , p = 12 3' (36) 2 1 ⎜ ⎟ 3' 2' 3' ⎝ Vm ⎠ m12 ⋅cv Vm γ−1 Q ⎛ VM ⎞ p m12 ⋅ r ⋅T4' T = T ⋅⎜ ⎟ (47) T4' = T3' + , V4' = (37) 2 1 ⎜ V ⎟ m12 ⋅cp p3' ⎝ m ⎠

W12 = m1 ⋅ (u2 − u1) (48) W3'4' = −p3' ⋅ (V4' − Vm ) (38) The conventional combustion cycle The expansion process from 4’ to 5’ is efficiency can be written as: considered adiabatic and reversible: ⎛m ⋅(u − u + u − u ) ⎞ γ−1 ⎜ 1 2 1 5' 4 ⎟ ⎛ V4' ⎞ ⎜ ⎟ ⎜ ⎟ − p3 ⋅()V4 − Vm + (pe − pi )⋅(VM − Vm ) T5' = T4' ⋅⎜ ⎟ (39) η = ⎝ ⎠ (49) ⎝ VM ⎠ c mf ⋅QLHV

W4'5' = m12 ⋅ (u5' − u4' ) (40) 3.6 Engine friction The pumping work of the low pressure To compute the effective torque produced cycle can be expressed as: by the engine from the indicated work, the friction losses of the engine must be evaluated:

Wp = ()pe − pi ⋅(VM − Vm ) (41) Wf = Wc − We (50) The supercharged hybrid pneumatic engine For this study a simple model is enough to efficiency can be written as the quotient of work perform the comparison between conventional produced (23, 29, 32, 38, 40, 41) and thermal and pneumatic hybrid engine operation. From the energy provided by the combustion (28): many friction models available, the Bidan model has been chosen for its simplicity (Chamaillard ⎛m ⋅()u − u − p ⋅(V − V ) ⎞ ⎜ 1 10 1 t 12 10 ⎟ et al., 2001). This black box model estimates ⎜+ m12 ⋅()u2' − u12 + u5' − u4' ⎟ friction losses in a global way. The only variable ⎜ ⎟ parameter is the engine rotational speed (N): ⎝− p3' ⋅()V4' − Vm + (pe − pi )⋅(VM − Vm )⎠ ηs = (42) 2 mf ⋅QLHV Wf = (VM − Vm )⋅(a + b⋅ N + c⋅ N ) (51) 3.4 Pneumatic undercharging Parameters a, b, c have been determined from engine measurements. As can be seen on Figure 6, the undercharged cycle only differs from the

Int.J. Applied Thermodynamics, Vol.5 (No.1) 7 3.7. High pressure tank controller. Thus, the global efficiency of the The air storage tank is considered adiabatic engine over the test cycle not only depends on and of constant volume. Therefore, the mass and the thermodynamic cycles and the thermo- energy balance can be written: physical properties of the fluid, but also on the test cycle itself and the engine control strategies. m = ∆m (52) t ∑ t The driving cycle consists of a vehicle speed versus time graph that must be followed by Where ∆m t represents the variation of the driver. The New European Driving Cycle mass inside the tank for one engine cycle and is equal to the variation of mass inside the cylinder (NEDC) (European Directive, 1990) has been during the tank charging and engine charging used for this study (Figure 7). processes (Equations 3, 12 and 25). 4.1 Vehicle and driveline model To determine the tank temperature, two To be able to follow the driving cycle, a distinct situations must be considered: dynamic vehicle model has to be used. It can be In the case of tank charging, the variation of expressed by: tank temperature during a single cycle can be Fw − Fr − Fa − Fb (55) written: x& = ∫ dt m v γ ⋅ T − T ∆T = c t ∆m (53) Where Fw represents the driving force at t m t t the wheels, Fr the rolling drag force, Fa the

In the case of tank discharging: aerodynamic drag force and Fb the braking force: ()γ −1 ⋅Tt ∆Tt = ∆m t (54) kd⋅ηd (56) m t Fw = ⋅ We 4 ⋅ π ⋅ rw The state of charge of the high pressure air Fr = m v ⋅g ⋅ k r (57) tank ( mt , Tt , p t ) is computed from equations 52, 53, 54 for each engine cycle. 2 ρ⋅A f ⋅C x ⋅ x& Fa = (58) 4. Driving Cycle Simulation 2

Unlike the efficiency of conventional Fb ≥ 0 (59) engines, the global efficiency of the hybrid The braking force only exists to close the pneumatic-combustion engine cannot simply be force balance in case of strong deceleration, computed from the thermodynamic cycles and when engine braking is not sufficient. the thermo-physical properties of the . In the case of a conventional engine simulation, the engine cannot produce torque at 0

120 rpm. Therefore, a clutch model must be added:

ωe > ωd (60) 100

We ⋅ ωe − ωd

) 80 P = (61)

h s / 4 m ⋅ π k

eed ( 60 p Where ωe and ωd respectively represent e s cl

hi the engine and driveline speeds, P the engine e s

V 40 power dissipated into heat during clutch slipping. 20 4.2 Operating mode selection rules

0 During the conventional internal 0 200 400 600 800 1000 1200 Time (s) combustion engine driving cycle simulation, the following set of rules has been adopted: Figure 7. New European Driving Cycle (NEDC) • If engine speed drops below 1000 rpm, During the hybrid engine operation over a downshift one gear. If engine speed goes beyond test cycle, operating modes must be switched to 2500 rpm, shift one gear up. achieve best global efficiency as well as tank refilling. The choice of the current operating • If engine speed drops below 1000 rpm in first gear, the clutch is slipping. mode is performed at any time by the engine

8 Int.J. Applied Thermodynamics, Vol.5 (No.1) During the hybrid pneumatic-combustion mv 800 kg engine driving cycle simulation, the following k 0.00863 set of rules has been used: r kd 13.45 7.57 5.01 3.77 2.84 • When the vehicle starts, use pneumatic r 0.282 m motor mode as long as the tank is not empty. w 0.8 Then switch to undercharge mode to fill up the ηd 1.293 kg/m3 tank to 80% of its capacity. ρa 2.06 m2 • If the tank is at 80% of its capacity and Af speed is greater than 40 km/h, use conventional Cx 0.312 combustion mode. volume as well as its initial state of charge can be • In braking situations, switch to pneumatic found in TABLE II and the vehicle simulation pump mode. parameters can be found in TABLE III. The vehicle data (aerodynamic parameters, gear • If full load conventional combustion mode ratios, wheel size) have been taken from a real, torque is not high enough, switch to small size car. supercharged mode. 5.2 Conventional operation results The global average efficiency can then be computed from the total fuel consumption and On Figure 8 it can be seen that the idle the total work produced by the powertrain. phases account for almost one-third of the total driving cycle time (373.8 seconds). The energy 5. Simulation results dissipated during this time could be saved if the 5.1 Model parameters engine was stopped. The engine simulation parameters can be found in TABLE I. The friction parameters of the engine have been fitted to a real two-liter engaged engine (Chamaillard et al., 2001). The air tank

slipping clutch TABLE I. ENGINE SIMULATION

PARAMETERS braking 3 VM − Vm 2000 cm idling

VM Vm 10

stop r 288 J/K⋅kg 0 200 400 600 800 1000 1200 Time (s) γ 1.4 Figure 8. Conventional engine operating modes pi 0.1 MPa

pe 0.11 MPa Each start comprises a phase with a slipping

Ti 300 K clutch. Again, this energy could be saved if the internal combustion engine could provide torque QLHV 42500 kJ/kg at very low speed. But this amount of energy is ηc 0.8 very low compared to the total energy consumed εp 0.5 (TABLE IV).

εv 0.5 a 0.776 TABLE IV. ENERGY BALANCE OF CONVENTIONAL ENGINE OPERATION b 0.189 c 0.0209 Total energy consumed 26.4 MJ Total mechanical energy provided 2.10 MJ Total braking energy 0.92 kJ TABLE II. AIR TANK PARAMETERS AND INITIAL VALUES Total clutch slipping energy 4.19 kJ 3 Global driving cycle efficiency 7.96% Vt 50 dm The slowdown phases being relatively pt 2 MPa smooth on the NEDC driving cycle, most of the T 300 K t slowdown can be performed with engine braking alone. The brakes are activated only once during TABLE III. VEHICLE SIMULATION the final slowdown of the extra-urban part of the PARAMETERS cycle (Figure 8). So, it seems that the pneumatic

Int.J. Applied Thermodynamics, Vol.5 (No.1) 9 pump mode would not be really useful on this kind of driving cycle (TABLE IV). 5.3 Hybrid operation results The hybrid engine operation displays a great improvement in global driving cycle efficiency compared to the conventional engine operation (TABLE IV). The fuel economy is improved by more than 15% (TABLE V). TABLE V. ENERGY BALANCE OF HYBRID ENGINE OPERATION Total energy consumed 22.4 MJ Total mechanical energy provided 2.10 MJ Braking energy recovered 0.46 kJ Figure 10: Air tank pressure Global driving cycle efficiency 9.38% On Figure 10, it must be noted that the final On Figure 9, it can be seen that with the tank pressure is the same as the initial pressure. engine chosen for this study, supercharged mode Thus, the energy balance is not biased by the air is never necessary. To further enhance the global tank contribution. Further optimization could be efficiency, the engine should be downsized to a achieved by selecting a more optimal tank one-liter displacement. The friction losses could volume as well as minimum and maximum tank then be reduced and the thermodynamic pressure thresholds. efficiency of the cycle could be further On Figure 11, the main differences between improved. conventional and hybrid operation are apparent. At each stop, the hybrid engine displays a plateau in the fuel mass consumed where the conventional engine is idling. When the hybrid undercharged engine is operating in combustion mode, the fuel consumption slope is steeper than in supercharged conventional mode. Indeed, the hybrid engine

conventional needs to refill the air tank each time it switches to combustion (undercharging) mode.

pump

0.7 motor Conventional operation 0.6 Hybrid operation

Stop 0 200 400 600 800 1000 1200 Time (s) 0.5 g) k

( d e Figure 9. Hybrid engine operating modes 0.4 onsum c

The amount of energy recovered during the ss 0.3 a slowdown phases is very small. This can be m attributed to the smooth slowdown of the driving Fuel 0.2 cycle: most of the time the friction losses of the engine are sufficient to slowdown the vehicle 0.1 considered. A heavier vehicle should allow more 0 0 200 400 600 800 1000 1200 energy to be recovered. Time (s)

Figure 11. Fuel mass consumed 6. Conclusion A new kind of hybrid pneumatic- combustion engine has been modeled using simple thermodynamic cycles. To be able to compare this engine to a conventional internal combustion engine, the simulation had to take into account the driving cycle, vehicle configuration as well as the mode switching strategy.

10 Int.J. Applied Thermodynamics, Vol.5 (No.1) The simulation of a transient driving cycle ζ : pneumatic pump efficiency shows a great improvement of the global ω : rotational speed efficiency with the new hybrid concept. The global fuel consumption could be reduced by Subscripts around 15% without any specific optimization. a : aerodynamic Regenerative braking is only applicable for b : brake quite harsh braking situations. In most usual c : combustion, cycle driving situations, the engine internal friction is d : driveline sufficient to dissipate the kinetic energy during a e : exhaust, engine deceleration phase. f : fuel, friction Further improvement could be achieved by i : intake using a downsized engine configuration. Tank LHV : lower heating value volume, tank pressure thresholds and mode M : maximum switching strategy optimization should also lead m : minimum to significant reduction of fuel consumption. p : at constant pressure r : rolling Nomenclature s : slipping Symbols t : tank v : at constant volume, vehicle Af : frontal area w : wheel a,b,c : friction coefficients c : specific heat Acronyms

Cx : drag coefficient DI: Direct Injection F : force TDC: Top Dead Center g : acceleration of gravity BDC: Bottom Dead Center h : enthalpy IMEP: Indicated Mean Effective Pressure k : gear ratio NEDC: New European Driving Cycle m : mass SOC: State Of Charge N : engine speed References P : power p : pressure Chamaillard, Y., Labreuche G., Da Costa, A., Charlet, A., Higelin, P., Perrier, C., 2001, “Total Q : heat Friction Effective Pressure and Torque r : gas constant, radius Estimator”, Modeling, Emissions and Control in T : temperature, torque Automotive Engines, (MECA’01). u : internal energy V : volume European Directive, 90/C81/O1, ECE Official W : work Journal C81, 1990. x : distance Higelin, P., Charlet, A., 2001, “Thermodynamic Greek symbols Cycles for a New Hybrid Pneumatic-Combustion Engine Concept”, 5th International Conference on γ : specific heat ratio Internal Combustion Engines (ICE 2001). ε : heat fraction η: efficiency Schechter, M., 1999, “New Cycles for Automobile Engines”, SAE paper -01-0623. ξ : pneumatic motor efficiency

Int.J. Applied Thermodynamics, Vol.5 (No.1) 11