FULL HYDRO case set up (PFI)
Aim: To simulate the full hydro setup of a Port Fuel Injected engine which uses Iso octane as the fuel and based on the results from the simulation calculate the following:
1.Compression Ratio
- Combustion Efficiency
- Power and Torque
- Significance of CA10, CA50 and CA90
Geometry and Boundary Flagging: The geometry and the boundary flagging aspects have been explained in detail on the previous project. The geometry was imported in Converge using the STL file given and the surface preparation and boundary flagging was done. Once the No-Hydro simulation ran smoothly and after the results obtained from it were verified, the full hydro case setup was done.
Case Setup: A crank angle-based IC engine application was chosen to perform the simulation. The engine dimensions (bore, stroke and connecting rod length) was provided as inputs. The engine speed was chosen to be 3000 RPM
The fuel chose was Iso Octane (iC8H18) and the spray parcel properties were taken from the data file of Iso Octane.
The transport parameters were set as default values. Reaction mechanism was set based on the data file for Iso octane combustion.
A full hydrodynamic solver is used in this case and the simulation was run for -520 Crank Angle Degree to 120 CAD with the initial time step being 1e-07 s and final time step being 0.0001s.
The solver parameters were taken as default values.
Boundary Conditions:
The piston boundary was set at a temperature of 450 K and the motion type was set as translating. The piston speed is 3000 RPM (as mentioned earlier). The cylinder head and liner were also assumed to be at the same temperature as that of the piston and they were set as stationary wall boundaries with Law of Wall condition.
Air was assumed to be the working fluid entering the intake port (inflow) at a temperature of 363K and at a pressure of 1 bar.
The intake ports were assumed to be at a temperature of 425 K and they were set as stationary boundaries with Law of Wall condition.
The intake valve boundaries were set at 480 K and they were set as translating wall boundaries. The intake valve lift input file was provided for specifying the valve lift at a given crank angle.
The exhaust ports were assumed to be at a slightly higher temperature (500K) since the gas flowing out of the exhaust ports will have a higher temperature.
Similarly, the exhaust valves were also set at higher temperatures (525 K) and were assumed to be translating boundaries with exhaust valve lift provided as input.
The exhaust gas flowing out of the outflow boundary is assumed to be at a temperature of 800 K and the stoichiometric composition of each compound contained in the exhaust gas was calculated based on the combustion reaction for Iso-Octane and then the mole fraction was given as input.
Initial Conditions: The initial conditions can be assigned to the setup based on four regions:
- Cylinder
- Intake Port-1
- Intake Port-2
- Exhaust Port
Cylinder Initial Conditions: The initial cylinder pressure is set as 1.85731 bar and the temperature as 1360 K. The temperature is set high because the in-cylinder temperature at the end of the combustion process will be extremely high. The composition of species contained inside the cylinder was computed using the stoichiometric proportions of the combustion reaction.
Intake Port: In the intake port near the combustion chamber (Intake Port-1) air is assumed to be flowing in with the fuel being injected through the nozzle. Thus, the mass fractions of oxygen, nitrogen are assumed to be air’s composition and the amount of fuel injected is also given. The temperature is assumed to be slightly higher than atmospheric temperature -390K and pressure is 1 atm. In the intake port-2 which is closer to the inflow boundary, the temperature is 370K and pressure 1 atm. Mass fraction is that of airs.
Exhaust Port: The temperature is assumed to be 500 K and pressure 1 atm. The composition of gas in the port is calculated by the stoichiometry of end products of combustion.
Events: The intake valve and exhaust valve opening and closing are controlled by an event. The valve lifts are provided as the input to control the opening and closing.
Spray Modelling: A Frossling spray model was used in this simulation to model the fuel spray. The parameters involved in the spray model was set to be the default values. The spray wall interaction model was chosen to be a wall-film one in the assumption that a film like layer would get created if the spray comes in contact with the wall. A Kelvin-Helmholtz injector model was used to model fuel injection. The total mass injected was set as 3 e-05 Kg and the rate shape value as set as below:
Four nozzles have been used in this simulation and their respective coordinates have been given as inputs. And the injection type is a solid-cone spray. An O’Rourke film splash was used.
Combustion Model: A SAGE model was used to model the combustion process in this simulation. The start time was specified as -17 CAD and end time as 130 CAD. The SAGE model was assumed to be region independent for simplicity.
An RNG k-epsilon turbulence model with default parameters was used for turbulence modelling.
The spark was modelled using a source model. The start of spark ignition is set as -15 CAD and the duration is 10 degree. The spark radius is set as 0.0005 m. The spark temperature is set as 50000 K.
The base grid was set around 1 mm since the target cell count was around 800000. AMR was used for the cylinder and intake ports to improve the accuracy. Embedding level was set as 3 and sub grid criterion of 2.5 was used.
A cyclic embedding for the injector was used as embedding needs to be active only when the fuel is sprayed. The entire piston cylinder arrangement was given an embedding at 2 levels in the scale of 1. The intake and exhaust valves were also provided with embedding.
RESULTS AND DISCUSSION:
The basic mesh creation is as shown below. The regions where finer mesh is seen were provided with embedding.
The pressure, temperature and volume graphs were obtained after the simulation was run as seen below:
The engine cycle in this case is assumed to be starting from the beginning of exhaust stroke. As the cylinder moves up pushing the exhaust gas out, the volume decreases and the pressure remains more or less constant at 1 bar. Similarly, when the intake stroke occurs the pressure is nearly atmospheric as atmospheric air is drawn into the cylinder. As the compression stroke occurs the volume reduces and pressure rises. The pressure reaches the peak at some point during the combustion process. This will be discussed in detail later. After the combustion, expansion occurs and the pressure drops.
Based on the results of the simulation we can calculate the following:
- Compression Ratio: The compression ratio of an IC engine is given by :
where Vd is the displacement volume and Vc is the clearance volume.
The maximum and the minimum cylinder volume can be found out from the graph as below:
Based on the above calculation the compression ratio of this engine was found out to be 10.07. It is an ideal compression ratio for a spark ignition engine as not a lot of pressure is need for combustion. For SI engines compression ratio is usually between 8-12. For CI engines the CR is between 12-20.
2. Heat transfer in IC engines is a very vital phenomena and it has a direct effect on the engine performance. In an SI engine, (like in this case) a spark ignites the compressed air fuel mixture and a flame kernel propagates throughout the system. This flame kernel gets quenched near the wall and thus heat transfer occurs between the flame and the wall. Inside the cylinder due to turbulence, swirling and tumbling, the accuracy of measurement of critical parameters get affected. Hence due to the two reasons listed above a wall heat transfer model is used rather than CFD to predict wall temperature.
3. The combustion efficiency of an IC engine is given by:
The entire amount of chemical energy possessed by the fuel is not converted during the combustion process. Some amount of fuel goes unburnt and that chemical energy is not utilised. The extent to which complete combustion occurs is measured by combustion efficiency. The mass of fuel injected is given as 3*10^-5. The lower heating value of iso octane is given as 44.6 MJ/kg. The mass trapped by the engine is found from the picture below. Mass trapped multiplied by the enthalpy of formation of the elements will give the heat released by combustion.
The combustion efficiency of this engine was calculated as below:
4. Torque: As per the formula given in the challenge, the torque can be found out as:
Power(P) = (2*π*N*T)/60
T= 60P/(2*π*N)
The power was obtained from the engine performance calculator and the torque calculation was made as below:
5.Significance of ca10, ca50 and ca90: Ca10, ca50 and ca90 represent the crank angle during various stages of combustion. Ca10 represents the stage where 10% of the combustion is completed. Similarly, ca50 and ca90 represent the 50% and 90% stages. Generally, the duration/difference between the ca10 and ca90 stages of combustion is taken as the combustion duration where ca10 is assumed to be start of combustion and ca90 the end. Ca50 is assumed to be the stage where the energy release in combustion is maximum as it is nearly during the midway stage of the combustion process.
Conclusion: Thus, the full hydro simulation of a PFI engine was conducted and the various efficiencies of the engine were calculated.
References:
- John B. Heywood (1988), Internal Combustion Engine Fundamentals, McGraw-Hill
- Plengsa-ard* and M. Kaewbumrung (2017), CFD modelling wall heat transfer inside a combustion chamber using ANSYS forte