Week 10: Project 1 - FULL HYDRO case set up (PFI)

Objective: To simulate the Full-hydro case setup of PFI.

           1. What is the compression ratio of the engine?
           2. Calculate the Combustion efficiency of the Engine?
           3. Determine Power and torque of engine by using an engine performance calculator?
           4. What is the significance of CA10, CA50, and CA90?
           5. Why do we need a wall heat transfer model? Why can't we predict wall temperature from CFD simulation?

Given: Geometry of PFI is given in .STL file.

The parameters which are necessarily required in order to run a complete full hydrodynamic simulation of PFI are given below:
1. Engine Geometric Parameter
2. Run Parameters
3. Simulation time parameter
4. Boundary condition
5. Initial condition
6. Injection parameter
7. Nozzle Position
8. Spark Ignition parameters.

Case Setup:

We already have successfully done surface preparation and Boundary flagging of PFI geometry.

Application:
Here, we are going to simulate a newly designed PFI engine so we will be proceeding by selecting IC Engine (Crank-angle based). 

The engine geometric parameter is given as below:

Materials:

In IC engine simulation we will be more interested in knowing reactions inside the combustion chamber after combustion takes place, how the fuel will be injected inside the cylinder during suction/ Intake stroke of the engine, what will be participating species or what will be the type of fuel we are going to inject inside the cylinder. That all will be done in this Materials section by selecting the proper type of fuel and respective parameters.

IC8H18 (Iso-Octane)- It has physical properties that are closer to the physical properties of GASOLINE fuel. That is the reason we chose Parcel species as "IC8H18".
Since we have selected IC8H18 as parcel species we will require it's thermodynamic data to involve more species in reaction mechanism contained in the file named "therm.dat".

Boundary Conditions:

Simulation Parameters:

We are going to perform full hydrodynamic simulation and simulation time parameter is set as:

Initial Condition and Events:

Region and Initialisation:

Complete PFI geometry has been bifurcated into volumetric regions that contain all boundaries that we have flagged.

Initialization of all regions will be done as:

Events:

Use the Events dialog box to create custom CLOSE (activate disconnect triangles), OPEN (deactivate disconnect triangles), or VALVE events between two or more regions at defined intervals. Because CONVERGE cannot properly execute a simulation if surface triangles from two boundaries intersect, using the Events section is the only way to discontinue (or allow) flow between two regions.

Note: If we select a VALVE event, you must include a valve profile file.
That's why both the intake and exhaust lift profile provided.
Opening and closure of Intake and Exhaust valves play a key role in the IC engine. In order to simulate valves events are useful which is created as:

Physical Models:

Following models, we have incorporated in this case setup in order to capture flow physics accurately inside the IC engine.

1. Spray Modelling- To define injectors, nozzles, and other spray-related parameters.
2. Combustion Modelling- To activate various combustion models.
3. Source/Sink Modelling- To define sources of energy, momentum, turbulence, porous media, or species.
4. Turbulence Modelling- To define the parameters and constants associated with the selected turbulence model.

Grid Control:

The base grid we will be working on is 4mm.

AMR: Both velocity and temperature-based grid refinement are performing on a cylinder and intake port(near the combustion chamber) region.

Velocity Based AMR:

Temperature Based AMR:

Fixed Embedding:

Fixed embedding is done at particular locations to capture accurate data.

Embedded regions are:

What is the compression ratio of the engine?

Compression Ratio:

In a reciprocating engine, the static compression ratio is the ratio between the volume of the cylinder and the combustion chamber when the piston is at the bottom of its stroke, and the volume of the combustion chamber when the piston is at the top of its stroke, It is therefore calculated by the formula

Where:

        $nc=\frac{1240.72}{444000}\cdot 3E-$=+,Maximum cylinder Volume. Included both swept volume and clearance volume. 

 = displacement volume. This is the volume inside the cylinder displaced by the piston from the beginning of the compression stroke to the end of the stroke.

 = clearance volume. This is the volume of the space in the cylinder left at the end of the compression stroke.

       can be estimated by the cylinder volume formula

Where:

= cylinder bore (diameter)

 = piston stroke length

Because of the complex shape of Vc, it is usually measured directly. This is often done by filling the cylinder with liquid and then measuring the volume of the used liquid.

Calculate the combustion efficiency of this engine:

Combustion efficiency:

In practice, the exhaust gas of an internal combustion engine contains incomplete combustion products (e.g., CO, H2, unburned hydrocarbons, soot) as well as complete combustion products (CO2, and H2O).
Under lean operating conditions the amounts of incomplete combustion products are small. Under fuel-rich operating conditions these amounts become more substantial since there is insufficient oxygen to complete combustion. Because a fraction of the fuel's chemical energy is not fully released inside the engine during the combustion process, it is useful to define a combustion efficiency.

The combustion efficiency η_c is defined as the ratio between the energy released by the burnt fuel and the theoretical energy content of the fuel mass during one complete engine cycle.

where:

H_R [J] – enthalpy (internal energy) of the reactant
H_P [J] – enthalpy (internal energy) of the product
T_A [K] – ambient temperature
m_f [kg] – a mass of fuel inducted per cycle
Q_HV [J/kg] – a heating value of the fuel

Integrated Heat Rate: It indicates the total amount of energy we are getting from the combustion process alone.

Combustion Efficiency will be calculated as:

From the graph,

Output power is 1240.72 J

Energy content of fuel=_{$\text{Mass of fuel inducted per cycle * Fuel LHV}$}

Where, Fuel Lower Heating Value (LHV) of ISOOCTANE is 44400 KJ/Kg
            Mass inducted is 3E-5

`n c = 1240.72 / 444000 * 3E-`

The combustion efficiency will be found as,

$\text{Combustion efficiency}\left(n_c\right)=\text{0.9315 or 93.15\%}$

$\text{Select Add Files}\to \text{Put 'thermo.out' file}$

And

$\text{Go to Load from engine.in}\to \text{Put 'engine.in' file}\to \text{Calculate}$

Torque Calculation is as:

Why do we need a wall heat transfer model? Why can't we predict wall temperature from CFD simulation?

$\text{Need of Wall Heat transfer Model}\to$

The Heat loss through the walls of the engine is an important parameter during engine optimization, as it influences power, efficiency and emissions, and engine component thermo-mechanical reliability. As a consequence, a correct estimation of wall heat fluxes is mandatory. However, due to the complexity of the phenomenon involved in the heat transfer mechanism, such estimation is far from being trivial. Hence, accurate modeling techniques need to be available.
An important aspect of the optimization procedure of these engine technologies is the study of heat transfer through the walls of the engine. this has a direct influence on the power and efficiency of the engine since more heat is being lost means less power being transferred to the crankshaft. Besides, this also influenced on the thermal energy that is going out through the exhaust, due to which different after treatment tools are required. It also influenced the formation of pollutants. In fact, gas temperature along with oxygen concentration is the main responsible for the Nitrogen oxide formation and they affect unburnt hydrocarbons and Soot as well. As legislation is imposing more and more stringent constraints on tailpipe emissions
in internal combustion engines, engine designers are pushing for alternative solutions. 

$\text{Prediction of Wall Heat Temperature from CFD simulation}\to$

As near-wall gas temperatures are relevantly higher than wall temperatures during compression and combustion, engine flows cannot be considered isothermal at all. As a consequence, models able to account for variations of gas properties with temperature were proposed.
We know N-S equations are governing equations of fluid dynamics and to solve it numerically we have multiple commercial codes based on FVM. In this method, the computational domain is being decomposed at a number of small finite volumes or computational cells.
As we know this process of decomposition is called meshing. The bodies in the flow are often treated as "VOID" or "HOLES". The surface of the body becomes one of the boundaries of the computational domain and the appropriate boundary conditions must be set on it to
obtain a numerical solution.

e.g. For incompressible flows, we have to solve only for momentum equations (i.e. equations for components of fluid velocity vector). For such a case treating a body as a “hole” in the mesh seems to be a reasonable approach – the body is solid, there is nothing to flow within it and in the body-connected frame of reference, all the velocities are zero! The wall boundary conditions, in this case, are “no-slip” (zero velocity at the surface) for viscous flow or “slip” (zero normal velocity) for inviscid flow.

A boundary condition for the energy equation at the surface of the body is ADIABATIC WALL CONDITION for most of the cases. Applying this boundary condition seems reasonable in case that we assume there is no heat transfer between flow and the body if we have perfect insulator on the surface of the body. Modern codes have the capability of solving energy
equations both for fluid flow and for a body (for a solid body it turns to heat transfer equation). Such a problem is called “conjugated heat transfer problem”

What is the significance of CA10, CA50, and CA90?

Mass Fraction Burn (MFB) is used to evaluate ignition delay and combustion duration by dividing the amount of accumulated fuel or accumulated heat during the combustion period by the total fuel or total heat release.
The combustion state occurring in the cylinder can be distinguished by MFB.

$\star$ From the start of fuel injection, the crank position where MFB becomes 10% is noted as CF10, the point where MFB becomes 90% is noted as CA90.

$\star$ CA50 denotes where the MFB is 50%, which means that 50% of injected fuel is converted into energy.

$\star$ The difference between fuel injection point start and CA10 is called Flame-development Angle or Ignition Delay. And the Difference between CA10 and CA90 is called Rapid-burning Angle or Combustion Duration

This engine configuration :

Results:

PLots of Parameters of the Combustion chamber:

PV-Diagram:

Here, it resembles with PV diagram of the OTTO cycle. Since we don't have some output files diagram doesn't have a closed-loop.

Temperature variation over the period of time is:

Here, we can see how flame propagation takes place inside the combustion chamber after initiating spark in the spark plug. Also, we can see how well AMR has performed inside the cylinder in order to capture the data at every crack angle.

Pressure and Temperature variation over the period of time is: