3. Modelling of physical processes in piston pneumatic engine

The mathematical model of the pneumatic engine was carried out to determine the engine performance at different control parameters. Calculation of the air mass delivered to the cylinder by determination of velocity and density of the air in inlet duct in front of the pneumatic valve enables assessment of engine work time at given tank volume and initial pressure. The air was treated as semi-perfect gas, where the specific heat ratio was calculated every time step [22].

## 3.1. Mass balance

2.2. Operation of pneumatic engine

132 Improvement Trends for Internal Combustion Engines

Figure 1. Non-isentropic work during air expansion.

The work performed by the pneumatic engine depends on the pressure difference between higher and lower heat source. The air expansion process is shown in Figure 1 from pressure p<sup>1</sup> to pressure p<sup>2</sup> with temperatures T<sup>1</sup> and T2, respectively [20]. The thermodynamic process between point 1 and point 2 is non-isentropic process, and the work l<sup>s</sup> has lower value than the isentropic process [21]. In order to obtain the higher power during one work cycle, the higher pressure of the higher heat source (tank) is required. If temperature of the air in the tank has value near the ambient temperature T<sup>1</sup> then temperature of the expanded air T<sup>2</sup> has lower temperature than ambient temperature. But, in a real pneumatic engine, the air injection takes place at the maximum value of compressed air delivered to the cylinder during the intake

process. Therefore, the temperature T<sup>1</sup> has higher value than ambient temperature.

The engine is filled only by the air at high pressure when the piston is at TDC. The pneumatic engine can be simply done by modification of the design of the classic two-stroke engine. The engine does not require the inlet port delivering the air to the crankcase. The crankcase has a vent that causes only small compression of the air. The oiling of the bearings and the cylinder surface is ensured by a small oil pump or by oil drop valve in a close cycle. The schematic idea of the pneumatic two-stroke engine and timing of valve and port opening are shown in Figure 2. Only one exhaust port is used for the gas exchange in the cylinder. The engine has an injector or pneumatic valve controlled by the electronic unit. The bottle of certain volume contains the air at high pressure. The pressure of stored air in the bottle or tank (about 300 bar) is reduced by a pressure regulator to smaller injection pressure about 20–30 bar. The pressure is controlled by the sensor and the air is delivered by the pipe of small diameter (about 5–8 mm) to the valve. The air volumetric flow rate through the valve is rather high in comparison to the liquid fuel injection. The use of the electromagnetic stem valve requires high voltage and high electric power. For that case, the electromagnetic pneumatic valve used in industry is better solution. The air flow control should enable the high pressure in the cylinder after top dead centre (ATDC), and on the other hand, the opening of the pneumatic valve lasts very short (about 40–60 CA) and due to this reason, the natural frequency of the moving elements

> The air thermodynamic parameters in the pipes and ducts were determined at assumption of unsteady gas flow from the three hyperbolic nonlinear partial differential equations: mass, momentum and energy balance. The system of the equations was solved by using the Lax-Harten-Leer scheme [23] based on Godunov's method. The engine parameters were

determined on the basis mass and energy balance of the charge in the cylinder. Based on the mass balance law, the increment of the air mass in the cylinder can be expressed by the following equation:

$$dm\_{\rm c} = F\_1 \cdot \mu\_1 \cdot \rho\_1 \cdot dt + F\_{\rm inj} \cdot \mu\_{\rm inj} \cdot \rho\_{\rm inj} \cdot dt - F\_2 \cdot \mu\_2 \cdot \rho\_2 \cdot dt \tag{1}$$

In most cases, the air injection takes place at a critical flow, because pressure in the injector is several times higher than in the cylinder. The critical flow occurs when the following condition

> 2 k þ 1 � � <sup>k</sup> k�1

Gaseous constant R of the air in the cylinder does not depend on temperature and k is the specific heat ratio and should be calculated for every considered time step Δt on the basis of

The mathematical model of gas flow in the pipes and engine ducts takes into account pressure wave motion, which means that flow is unsteady. Non-dimensional velocity A = u/â of the air flown into the cylinder is calculated on the thermodynamic equations for isentropic unsteady

> A2 c � �<sup>c</sup>þ<sup>1</sup>

> > <sup>B</sup><sup>2</sup> � <sup>1</sup>

Gas velocity u is calculated from the given nonlinear equations by solving variable A. The same equations enable the calculation of the air outflow velocity in the pipe near the exhaust port.

The amount of the heat transfer to the walls is calculated on the basis of the conductive heat coefficient hc, area of heat exchange Fh and the temperature difference between gas Tc and

¼ 0 (7)

Modern Pneumatic and Combustion Hybrid Engines http://dx.doi.org/10.5772/intechopen.69689

<sup>B</sup><sup>2</sup><sup>c</sup> � <sup>Ψ</sup><sup>2</sup> <sup>Ψ</sup><sup>2</sup> (8)

<sup>k</sup>þ1, â = substitute of gas sound speed, p^ = substitute of pressure, Ψ

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2 k þ 1 � �<sup>k</sup>þ<sup>1</sup> k�1

s

(5)

135

(6)

pc pinj ≤

dt <sup>¼</sup> <sup>Φ</sup> � Finj � pinj � <sup>k</sup> ainj

the change of temperature in the cylinder and pipes [23] which means that k=f(T).

<sup>b</sup><sup>c</sup> � b B<sup>2</sup> <sup>þ</sup>

<sup>A</sup><sup>2</sup> <sup>¼</sup> <sup>c</sup>

In such a case, the mass flow rate is calculated as follows:

3.5. Unsteady gas flow in feeding ducts

gas flow through contraction:

a. For sonic flow:

b. For subsonic flow

where B ¼ <sup>p</sup>

walls Tw [24]:

<sup>p</sup>^, c ¼ <sup>2</sup>

= general flow coefficient.

<sup>k</sup>�1, b ¼ <sup>2</sup>

3.6. Heat transfer and kinematic dependencies

dminj

A2 B<sup>2</sup><sup>c</sup> Ψ<sup>2</sup>

is fulfilled:

where F<sup>1</sup> = cross section area of the inflow pipe (delivered air from the regulator), F2 = cross section area of the outflow port, Finj = cross section area of injector nozzle, u = charge velocity in the pipes, ρ = charge density in the pipes, t = time, inj = air injector.

#### 3.2. Conservation of energy

Change of the internal energy U in the time increment dt is determined by the formula:

$$d\mathcal{U} = \mathbf{i}\_1 \cdot dm\_1 + \mathbf{i}\_{\text{in}\mathfrak{j}} \cdot dm\_{\text{in}\mathfrak{j}} - \mathbf{i}\_2 \cdot dm\_2 + d\mathbb{Q}\_h - p\_c dV \tag{2}$$

where Qh = heat exchange with walls, i = enthalpy of the air at defined temperature, V = cylinder volume, dm = mass flow rate in the pipes and through the injector, pc = cylinder pressure.

#### 3.3. Determination of cylinder pressure

After some simplifications and assuming k as specific heats ratio (k = cp / cν), the energy equation gives the formula of the pressure increment in the cylinder:

$$dp\_c = \frac{k-1}{V} \left( dQ\_h - kp\_c dV + kR(T\_1 dm\_1 + T\_{\text{inj}} dm\_{\text{inj}} - T\_2 dm\_2) \right) \tag{3}$$

Equation (3) does not contain the component of fuel combustion as in the real combustion engine. All these components depend on time, an increment of the inflow air mass takes place during opening of transfer ports and increment of outflow air mass takes place only during opening of the exhaust port, whereas the air injection lasts very short and begins when piston is near TDC.

#### 3.4. Mass flow rate through injector nozzle

The mass flow rate of the injected air depends on pressure difference between the injector and the cylinder. During calculation, one should check whether the flow is critical or subcritical. For the second case, the mass flow rate is determined from the following equation:

$$\frac{d m\_{\rm inj}}{dt} = \frac{\Phi \cdot F\_{\rm inj} \cdot p\_{\rm inj} \cdot \sqrt{k}}{a\_{\rm inj}} \sqrt{\frac{2k}{k-1} \left[ \left( \frac{p\_c}{p\_{\rm inj}} \right)^{\frac{2}{k}} - \left( \frac{p\_c}{p\_{\rm inj}} \right)^{\frac{k+1}{k}} \right]} \tag{4}$$

where Φ = flow resistance through the injector nozzle, Finj = injector flow area, pinj = pressure of injected air, <sup>a</sup>inj = sound velocity of injected air <sup>ð</sup>ainj <sup>¼</sup> ffiffiffiffiffiffiffiffiffiffiffiffiffi kRTinj <sup>p</sup> <sup>Þ</sup>, pc = pressure of air in cylinder and R is an individual constant of air.

In most cases, the air injection takes place at a critical flow, because pressure in the injector is several times higher than in the cylinder. The critical flow occurs when the following condition is fulfilled:

$$\frac{p\_c}{p\_{inj}} \le \left(\frac{2}{k+1}\right)^{\frac{k}{k-1}}\tag{5}$$

In such a case, the mass flow rate is calculated as follows:

$$\frac{d m\_{\rm inj}}{dt} = \frac{\Phi \cdot F\_{\rm inj} \cdot p\_{\rm inj} \cdot k}{a\_{\rm inj}} \sqrt{\left(\frac{2}{k+1}\right)^{\frac{k+1}{k-1}}} \tag{6}$$

Gaseous constant R of the air in the cylinder does not depend on temperature and k is the specific heat ratio and should be calculated for every considered time step Δt on the basis of the change of temperature in the cylinder and pipes [23] which means that k=f(T).

#### 3.5. Unsteady gas flow in feeding ducts

The mathematical model of gas flow in the pipes and engine ducts takes into account pressure wave motion, which means that flow is unsteady. Non-dimensional velocity A = u/â of the air flown into the cylinder is calculated on the thermodynamic equations for isentropic unsteady gas flow through contraction:

a. For sonic flow:

determined on the basis mass and energy balance of the charge in the cylinder. Based on the mass balance law, the increment of the air mass in the cylinder can be expressed by the

where F<sup>1</sup> = cross section area of the inflow pipe (delivered air from the regulator), F2 = cross section area of the outflow port, Finj = cross section area of injector nozzle, u = charge velocity in

where Qh = heat exchange with walls, i = enthalpy of the air at defined temperature, V = cylinder volume, dm = mass flow rate in the pipes and through the injector, pc = cylinder pressure.

After some simplifications and assuming k as specific heats ratio (k = cp / cν), the energy

Equation (3) does not contain the component of fuel combustion as in the real combustion engine. All these components depend on time, an increment of the inflow air mass takes place during opening of transfer ports and increment of outflow air mass takes place only during opening of the exhaust port, whereas the air injection lasts very short and begins when piston

The mass flow rate of the injected air depends on pressure difference between the injector and the cylinder. During calculation, one should check whether the flow is critical or subcritical.

> 2k k � 1

where Φ = flow resistance through the injector nozzle, Finj = injector flow area, pinj = pressure of

2 4

For the second case, the mass flow rate is determined from the following equation:

k p

dt <sup>¼</sup> <sup>Φ</sup> � Finj � pinj � ffiffi

injected air, <sup>a</sup>inj = sound velocity of injected air <sup>ð</sup>ainj <sup>¼</sup> ffiffiffiffiffiffiffiffiffiffiffiffiffi

ainj

Change of the internal energy U in the time increment dt is determined by the formula:

the pipes, ρ = charge density in the pipes, t = time, inj = air injector.

equation gives the formula of the pressure increment in the cylinder:

dmc ¼ F<sup>1</sup> � u<sup>1</sup> � ρ<sup>1</sup> � dt þ Finj � uinj � ρinj � dt � F<sup>2</sup> � u<sup>2</sup> � ρ<sup>2</sup> � dt (1)

dU ¼ i<sup>1</sup> � dm<sup>1</sup> þ iinj � dminj � i<sup>2</sup> � dm<sup>2</sup> þ dQh � pcdV (2)

dQh � kpcdV þ kRðT1dm<sup>1</sup> þ Tinjdminj � T2dm2Þ (3)

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi

kRTinj

 !<sup>k</sup>þ<sup>1</sup> k

3 5 vuuut (4)

<sup>p</sup> <sup>Þ</sup>, pc = pressure of air in cylinder

pc pinj

 !<sup>2</sup> k � pc pinj

following equation:

134 Improvement Trends for Internal Combustion Engines

3.2. Conservation of energy

is near TDC.

3.3. Determination of cylinder pressure

dpc <sup>¼</sup> <sup>k</sup> � <sup>1</sup> V �

3.4. Mass flow rate through injector nozzle

dminj

and R is an individual constant of air.

$$\frac{A^2 B^{2c}}{\Psi^2 b^c} - b \left(B^2 + \frac{A^2}{c}\right)^{c+1} = 0\tag{7}$$

b. For subsonic flow

$$A^2 = c \frac{B^2 - 1}{B^{2c} - \Psi^2} \Psi^2 \tag{8}$$

where B ¼ <sup>p</sup> <sup>p</sup>^, c ¼ <sup>2</sup> <sup>k</sup>�1, b ¼ <sup>2</sup> <sup>k</sup>þ1, â = substitute of gas sound speed, p^ = substitute of pressure, Ψ = general flow coefficient.

Gas velocity u is calculated from the given nonlinear equations by solving variable A. The same equations enable the calculation of the air outflow velocity in the pipe near the exhaust port.

#### 3.6. Heat transfer and kinematic dependencies

The amount of the heat transfer to the walls is calculated on the basis of the conductive heat coefficient hc, area of heat exchange Fh and the temperature difference between gas Tc and walls Tw [24]:

$$dQ\_h = -h\_c \cdot F\_h \cdot (T\_c - T\_w) \cdot dt \tag{9}$$

After finding cylinder pressure pc and knowing the charge mass mc, we can find temperature from the equation of general gas state:

$$T = \frac{p \cdot V}{m \cdot R} \tag{10}$$

Cylinder volume V is determined from the dependency:

$$V = \frac{\varepsilon}{\varepsilon - 1} V\_s \tag{11}$$

The filling pressure of 60 bar causes the pressure increase in the cylinder to maximum value 20 bar. The calculations were carried out at a valve opening 5 CA before top dead centre (BTDC)

Modern Pneumatic and Combustion Hybrid Engines http://dx.doi.org/10.5772/intechopen.69689 137

The higher indicated mean effective pressure (imep) value is obtained at lower compression pressure, which takes place at lower compression ratio. The air temperature inside the cylinder depends also on the filling pressure. Higher filling pressure causes higher temperature of the air in the cylinder. Variation of the cylinder temperature is presented in Figure 4. The calculations were performed at the same control parameters as for pressure calculations. One can observe very low temperature at the end of the expansion process. At low filling pressure, for

At higher injection pressure, temperature in the cylinder decreases and at value of 60 bar, the maximum temperature at TDC is below 400 K. This situation causes the transfer of heat from the walls to the charge in the cylinder. The characteristic of engine effective power has quite different variation than characteristic of the classic two-stroke engine (Figure 5). The engine has bigger power at low rotational speed at the same valve timing. The characteristic was obtained for air injection pressure of 25 bar. In the pneumatic two-stroke engine, the highest value of brake mean effective pressure (bmep) takes place at lowest rotational speeds. This phenomenon is like as in electrical engines. Higher bmep value at higher rotational speeds can be assured by higher filling pressure, which causes a bigger air dose injected by the valve to the cylinder. Another way of increasing of bmep is increasing the duration of valve opening at the same filling pressure. The increase of the air injection pressure causes almost linear increase of bmep (Figure 6), but this causes increase of specific air consumption (SAC) value. The

4.2. Thermodynamic parameters of pneumatic two-stroke engine

Figure 3. Cylinder pressure in a function of crank angle at different air injection pressure.

example 20 bar, the cylinder temperature decreases below 200 K.

and duration 40 CA.

where ε = compression ratio, Vs = engine piston displacement.

Cylinder piston displacement is determined from kinematics dependencies of the crank-piston system in dependence on crank angle position α:

$$V\_s = \frac{\pi \cdot D^2}{8} \cdot s \cdot \left(1 + \frac{\delta}{4} - \cos \alpha - \frac{\delta}{4} \cos 2\alpha\right) \tag{12}$$

where D = piston diameter, s = piston stroke, δ = crank ratio (δ = s/(2L)) and L = connecting rod length. The processes taking place in the cylinder, inlet valve and exhaust pipe are fully described in the literature [21, 22, 25]. The whole model takes into account a wave pressure motion in the pipes and changes of the thermodynamic parameters in each time step (semiperfect gas). The model enables the calculation of the pressure, temperature, density, air velocity in the inlet and outlet pipes and also the air consumption.

#### 4. Calculation results of two-stroke pneumatic engine

The presented mathematical model was the basis for the development of a computer program in order to simulate the processes taking place in the virtual pneumatic engine. Several works concerning to a piston pneumatic engine were published by Mitianiec and Wiatrak in engine's literature [26, 27]. For this elaboration, the calculations were carried out for different rotational speeds, different filling pressure of the air and valve control parameters.

#### 4.1. Geometrical parameters of engine and boundary conditions

The simulation process considers to the two-stroke engine Robin EC12 with bore D = 75 mm, stroke S = 55 mm and compression ratio 6.5 and opening of exhaust port at 106� CA ATDC, which was fully tested in the standard version. Initial pressure in the tank was assumed as 300 bar. The pipe connecting the reducer and the valve with length 80 mm amounted had 8 mm of diameter. The valve lift during opening was assumed as sinusoidal. The air temperature in the tank was near ambient temperature and amounted 300 K. The higher filling pressure also causes higher cylinder pressure, as is shown in Figure 3, at a rotational speed of 2400 rpm.

Figure 3. Cylinder pressure in a function of crank angle at different air injection pressure.

dQh ¼ �hc � Fh � ðTc � TwÞ � dt (9)

<sup>m</sup> � <sup>R</sup> (10)

Vs (11)

(12)

After finding cylinder pressure pc and knowing the charge mass mc, we can find temperature

<sup>T</sup> <sup>¼</sup> <sup>p</sup> � <sup>V</sup>

<sup>V</sup> <sup>¼</sup> <sup>ε</sup> ε � 1

Cylinder piston displacement is determined from kinematics dependencies of the crank-piston

δ

where D = piston diameter, s = piston stroke, δ = crank ratio (δ = s/(2L)) and L = connecting rod length. The processes taking place in the cylinder, inlet valve and exhaust pipe are fully described in the literature [21, 22, 25]. The whole model takes into account a wave pressure motion in the pipes and changes of the thermodynamic parameters in each time step (semiperfect gas). The model enables the calculation of the pressure, temperature, density, air

The presented mathematical model was the basis for the development of a computer program in order to simulate the processes taking place in the virtual pneumatic engine. Several works concerning to a piston pneumatic engine were published by Mitianiec and Wiatrak in engine's literature [26, 27]. For this elaboration, the calculations were carried out for different rotational

The simulation process considers to the two-stroke engine Robin EC12 with bore D = 75 mm, stroke S = 55 mm and compression ratio 6.5 and opening of exhaust port at 106� CA ATDC, which was fully tested in the standard version. Initial pressure in the tank was assumed as 300 bar. The pipe connecting the reducer and the valve with length 80 mm amounted had 8 mm of diameter. The valve lift during opening was assumed as sinusoidal. The air temperature in the tank was near ambient temperature and amounted 300 K. The higher filling pressure also causes higher cylinder pressure, as is shown in Figure 3, at a rotational speed of 2400 rpm.

<sup>4</sup> � cos <sup>α</sup> � <sup>δ</sup>

<sup>4</sup> cos 2<sup>α</sup>

from the equation of general gas state:

136 Improvement Trends for Internal Combustion Engines

Cylinder volume V is determined from the dependency:

system in dependence on crank angle position α:

where ε = compression ratio, Vs = engine piston displacement.

Vs <sup>¼</sup> <sup>π</sup> � <sup>D</sup><sup>2</sup>

velocity in the inlet and outlet pipes and also the air consumption.

4. Calculation results of two-stroke pneumatic engine

speeds, different filling pressure of the air and valve control parameters.

4.1. Geometrical parameters of engine and boundary conditions

<sup>8</sup> � <sup>s</sup> � <sup>1</sup> <sup>þ</sup>

The filling pressure of 60 bar causes the pressure increase in the cylinder to maximum value 20 bar. The calculations were carried out at a valve opening 5 CA before top dead centre (BTDC) and duration 40 CA.

#### 4.2. Thermodynamic parameters of pneumatic two-stroke engine

The higher indicated mean effective pressure (imep) value is obtained at lower compression pressure, which takes place at lower compression ratio. The air temperature inside the cylinder depends also on the filling pressure. Higher filling pressure causes higher temperature of the air in the cylinder. Variation of the cylinder temperature is presented in Figure 4. The calculations were performed at the same control parameters as for pressure calculations. One can observe very low temperature at the end of the expansion process. At low filling pressure, for example 20 bar, the cylinder temperature decreases below 200 K.

At higher injection pressure, temperature in the cylinder decreases and at value of 60 bar, the maximum temperature at TDC is below 400 K. This situation causes the transfer of heat from the walls to the charge in the cylinder. The characteristic of engine effective power has quite different variation than characteristic of the classic two-stroke engine (Figure 5). The engine has bigger power at low rotational speed at the same valve timing. The characteristic was obtained for air injection pressure of 25 bar. In the pneumatic two-stroke engine, the highest value of brake mean effective pressure (bmep) takes place at lowest rotational speeds. This phenomenon is like as in electrical engines. Higher bmep value at higher rotational speeds can be assured by higher filling pressure, which causes a bigger air dose injected by the valve to the cylinder. Another way of increasing of bmep is increasing the duration of valve opening at the same filling pressure. The increase of the air injection pressure causes almost linear increase of bmep (Figure 6), but this causes increase of specific air consumption (SAC) value. The

Figure 4. Cylinder temperature in a function of crank angle at different air injection pressure.

4.3. Air consumption in pneumatic two-stroke engine

Figure 6. Torque and specific air consumption in a function of injection pressure.

change rapidly for considered volume of the bottle.

that volume.

As internal combustion engines, the efficiency of the pneumatic engine can be determined by an amount of air mass needed for producing power unit. During valve opening, the air mass delivered to the cylinder was calculated as a sum of partial masses at every time step. Variations of SAC and air mass per cycle (AMPC) as a function of engine rotational speeds are presented in Figure 7 at an injection pressure of 25 bar. With increasing of rotational speed at the same angle of opening of the injector (shorter time), one can observe a decrease of air mass

Modern Pneumatic and Combustion Hybrid Engines http://dx.doi.org/10.5772/intechopen.69689 139

For higher specific air consumption, the total efficiency is lower at higher rotational speed. This indicates to use the pneumatic engine at lower rotational speeds. The simulation showed the dependence of emptying of the tank on the air injection pressure. Variation of time emptying in a function of the filling pressure is shown in Figure 8 at a rotational speed of 2400 rpm. Emptying time of the tank at air injection pressure of 20 bars for 55 minutes for the tank with volume 170 l and emptying time of tank with another volume will be almost proportional to

The simulation shows a small change of the emptying time at medium rotational speeds and constant air injection pressure. It is caused by influence of non-steady gas flow on the ducts of the two-stroke engine. This time is almost proportional to the tank volume. Figure 9 shows the variation of emptying time as a function of the engine rotational speed at air injection pressure of 25 bar of a tank with volume 100 l at full engine load (WOT). Emptying time does not

consumption per cycle, but the amount of cycles increases in the same period.

Figure 5. Effective power and torque in a function of rotational speed.

calculations were carried out for valve opening 5 CA BTDC, n = 2400 rpm, wide opening throttle (WOT) and the opening duration 40 CA.

The change of the engine torque requires an automatic control of the filling pressure in the reducer. The same graph shows variation of SAC, which depends linearly on the injection pressure. Lower specific SAC values occur at lower pressure of air injection. Thus very important is reduction of air pressure from high value in the bottle to lower value in the injector.

Figure 6. Torque and specific air consumption in a function of injection pressure.

#### 4.3. Air consumption in pneumatic two-stroke engine

calculations were carried out for valve opening 5 CA BTDC, n = 2400 rpm, wide opening

The change of the engine torque requires an automatic control of the filling pressure in the reducer. The same graph shows variation of SAC, which depends linearly on the injection pressure. Lower specific SAC values occur at lower pressure of air injection. Thus very important is reduction of air pressure from high value in the bottle to lower value in the injector.

throttle (WOT) and the opening duration 40 CA.

Figure 5. Effective power and torque in a function of rotational speed.

Figure 4. Cylinder temperature in a function of crank angle at different air injection pressure.

138 Improvement Trends for Internal Combustion Engines

As internal combustion engines, the efficiency of the pneumatic engine can be determined by an amount of air mass needed for producing power unit. During valve opening, the air mass delivered to the cylinder was calculated as a sum of partial masses at every time step. Variations of SAC and air mass per cycle (AMPC) as a function of engine rotational speeds are presented in Figure 7 at an injection pressure of 25 bar. With increasing of rotational speed at the same angle of opening of the injector (shorter time), one can observe a decrease of air mass consumption per cycle, but the amount of cycles increases in the same period.

For higher specific air consumption, the total efficiency is lower at higher rotational speed. This indicates to use the pneumatic engine at lower rotational speeds. The simulation showed the dependence of emptying of the tank on the air injection pressure. Variation of time emptying in a function of the filling pressure is shown in Figure 8 at a rotational speed of 2400 rpm. Emptying time of the tank at air injection pressure of 20 bars for 55 minutes for the tank with volume 170 l and emptying time of tank with another volume will be almost proportional to that volume.

The simulation shows a small change of the emptying time at medium rotational speeds and constant air injection pressure. It is caused by influence of non-steady gas flow on the ducts of the two-stroke engine. This time is almost proportional to the tank volume. Figure 9 shows the variation of emptying time as a function of the engine rotational speed at air injection pressure of 25 bar of a tank with volume 100 l at full engine load (WOT). Emptying time does not change rapidly for considered volume of the bottle.

Figure 7. Consumption of air mass per cycle and specific air consumption in a function of rotational speed at air injection pressure 25 bar.

5. Assessment of two-stroke pneumatic engine

(WOT).

basis of the simulation results, the following remarks can be drawn:

This air comes from the process of normal filling of the cylinder.

pressure to about 700 bar.

The numerical analysis of the work of the pneumatic two-stroke engine based on the mathematical model and results from the simulation was carried out by theoretical considerations and results obtained from computer program. Most of previously done research works presented in introduction and this work was concentrated on air two-stroke engines. On the

Figure 9. Emptying time of a tank with volume 100 l in a function of engine speed at air pressure 25 bar and full load

Modern Pneumatic and Combustion Hybrid Engines http://dx.doi.org/10.5772/intechopen.69689 141

1. The two-stroke engine with air injection indicates smoother work than the four-stroke engine for the same air injection pressure and enables better utilization of the compressed air than the rotor engine. Every rotation of crankshaft produces power and such engine only wasted the air to the exhaust port during the scavenge process, not during air injection. For the same rotational speed and pressure of air injection below 150 bar, the four-stroke engine indicates lower bmep and also lower SAC than the two-stroke engine.

2. Every pneumatic engine indicates higher value of bmep at lower rotational speeds and enables better starting of vehicle. Good characteristic of the two-stroke engine depends on very short

3. The time of engine work depends on the air filling pressure and tank volume. It should be emphasized short time operation of the pneumatic engine under heavy loads. It would add the tanks of greater volume or more resistant (tanks made from composites) to greater

crank angle injection with value near 40 CA and start of air injection about 5 CA.

Figure 8. Emptying time of air for a tank with volume 170 l at n = 2400 rpm and different injection pressure and full load.

The pneumatic valve controlled by the electronic unit must enable an adequate air mass flow rate in a short time. The dose of air per cycle is one of the most important parameters needed for the design of the pneumatic valve.

Figure 9. Emptying time of a tank with volume 100 l in a function of engine speed at air pressure 25 bar and full load (WOT).

## 5. Assessment of two-stroke pneumatic engine

The pneumatic valve controlled by the electronic unit must enable an adequate air mass flow rate in a short time. The dose of air per cycle is one of the most important parameters needed

Figure 8. Emptying time of air for a tank with volume 170 l at n = 2400 rpm and different injection pressure and full load.

Figure 7. Consumption of air mass per cycle and specific air consumption in a function of rotational speed at air injection

for the design of the pneumatic valve.

pressure 25 bar.

140 Improvement Trends for Internal Combustion Engines

The numerical analysis of the work of the pneumatic two-stroke engine based on the mathematical model and results from the simulation was carried out by theoretical considerations and results obtained from computer program. Most of previously done research works presented in introduction and this work was concentrated on air two-stroke engines. On the basis of the simulation results, the following remarks can be drawn:


The design of the pneumatic engine is based on the classic two-stroke engine and it needs small changes in order to mount the air feeding system with electronic control unit (ECU) system.
