1-D Thermodynamic Analysis for Tractor applications Engine using GT-SUITE

OBJECTIVE:

• A complete thermodynamic analysis for a tractor application engine with basic target values provided
• The power produced by an engine must be 75HP i.e. 55.92 KW
• BSFC target limits provided are Min:210 g/Kw-hr and Max:225 g/Kw-hr
• A cooled EGR configuration to be implemented and integrated with basic engine model

A random Indian tractor engine model was looked up to for reference.

A turbocharged engine is selected over a naturally aspirated engine as turbocharger helps to force a greater amount of air into the combustion chamber than a naturally-aspirated engine which depends on void/ vacuum created by in-cylinder. A turbocharger also helps with the increase in efficiency as well as an increase in Power Out as high torque is an important parameter for an application of heavy-duty carried out by tractors. The engine size is taken to be 3.7 Litres. Usage of the turbo-charger gives me an allowance of using a small size engine (i.e. higher CR) increasing thermal efficiency and mass flow rate of air and fuel, producing higher Mean Effective Pressure with smaller engine size when compared with a naturally aspirated engine.

Engine Model Geometry:

Attributes	Values
Bore	100 mm
Stroke	128.12 mm
Connecting Rod	256 mm
Compression Ratio	16.5
TDC clearance height	0.5

The engine geometry is modelled as seen in the figure below. It is a simple turbocharged 4 valve, 4 in-cylinder engine with an EGR system implemented as seen in the upper part of the geometry. The firing order used is 1- 3 - 4 - 2 which is generally used in Inline 4 cylinder engines.

The output of a compressor is been used as the inlet end environment along with the turbine inlet at the end for our model. The inlet environment is followed by a cooler with imposed wall temperature of 350K. The two sensor connections are attached to achieve and maintain the required amount of inlet air temperature with the help of the actuator as cooling the inlet air decreases its's density, allowing more air to enter the combustion chamber. This exit end of the cooler is attached with the flow split general where the cooled EGR gets mixed with the incoming air and further head towards the manifolds of the cylinders. As the cylinder is a 2 inlet valve and 2 exit valve, a flow split y template is used for the intake ports and the exhaust ports attached with the cylinder. Again two flow split general are used each for two combustion chamber exhaust which leads to flow split general with two outlets, one as an EGR inlet and other as a turbine inlet.

The amount of EGR is controlled by adjusting the EGR valve diameter. Usage of cooled EGR helps us to reduce the temperature of the exhaust gas diluent which further when enters the combustion chamber reduces the mean in-cylinder temperature reducing the NOx emissions formation.

The in-cylinder boundary conditions are given as follows:

Absolute Pressure: 1bar

Temperature: 300k

Composition (fraction): n2 - 0.767, o2 - 0.233

Now, this 1-D simulation is run for two points, namely, Rated Speed and Intermediate Speed. We have considered the rated speed as 2200 rpm and intermediate speed as 1300 rpm. And the A/F to be considered are 25 to achieve maximum power and 22 to achieve maximum torque.

We tried to achieve maximum power and maximum torque at 2200 rpm and 1300 rpm respectively. We calculated and wrote down the targets for speed, power, torque, bsfc, amount of fuel, Air-to-fuel ratio and airflow rate.

A closer look at the EGR system used

Targets to be achieved

Speed	2200 rpm	1300 rpm
Power	55.92 KW	55.92 KW
Torque	242.72 Nm	410.76 Nm

bsfc (g/Kw-hr)	225	225
fuel	12.582 kg/hr	12.582 kg/hr
	3495 kg/s	3495 kg/s
	36.66 mg/s	21.66 mg/s
	95.31 rps	161.3 rps
	190.63 mg/cycle	322.61 mg/cycle
	47.65 mg/cycle/cylinder	80.65mg/cycle/cylinder
A/F ratio	25	22
Airflow rate	314.55	276.804

Comparison between Target and the achieved simulation results:

1D CFD simulation was run to achieve the targets given and below are the comparison plots between the target value and predicted value. The simulations were run with the intent to achieve or reach the target values as close as possible i.e. minimum error. However, anything within a 10% error is considered to be good.

Figure 1 displays the power output comparison for 2200 rpm and 1300 rpm. For 2200 rpm i.e. Rated speed, CFD slightly under predicts the target value with an error of 2.41% whereas for 1300 rpm i.e. Intermediate speed, CFD slightly over predicts the target value with an error of 1.72%.

Figure 1: Power for 2200 rpm and 1300 rpm

Figure 2 displays the torque comparison for 2200 rpm and 1300 rpm. For 2200 rpm i.e. Rated speed, CFD slightly under predicts the target value with an error of 2.37% whereas for 1300 rpm i.e. Intermediate speed, CFD slightly over predicts the target value with an error of 1.72%.

Figure 2: Torque for 2200 rpm and 1300 rpm

Figure 3 displays the Brake Specific Fuel Consumption comparison for 2200 rpm and 1300 rpm. For 2200 rpm i.e. Rated speed, CFD slightly under predicts the target value with an error of 6.63% whereas for 1300 rpm i.e. Intermediate speed, CFD slightly over predicts the target value with an error of 1.30%.

Figure 3: BSFC for 2200 rpm and 1300 rpm

Figure 4 displays the Airflow rate comparison for 2200 rpm and 1300 rpm. For 2200 rpm i.e. Rated speed, CFD slightly under predicts the target value with an error of 5.7% whereas for 1300 rpm i.e. Intermediate speed, CFD slightly over predicts the target value with an error of 0.17%.

Figure 4: Airflow rate for 2200 rpm and 1300 rpm

PLOTS

The comparison was performed for some important parameters between 2200 rpm and 1300 rpm. But we cannot make a direct comparison as compressor out pressure (absolute) for 2200 rpm and 1300 rpm are 1.5 bar and 2.33 bar respectively. To achieve the target power of 75HP, compressor out pressure customization was required.

Below is the In-cylinder pressure - theta plot comparison for 2200 rpm and 1300 rpm, where we observe pressure plot reaching a peak value at 5.68 Crank Angle deg. for both the cases. For 2200 rpm, peak pressure achieved is 66 bar whereas, for 1300 rpm, it reaches up to 92.4 bar. The pressure achieved for 1300 rpm is approximately 40% higher than 2200 rpm.

Figure 5: Pressure plot for 2200 rpm and 1300 rpm

Figure 6: Pressure-Volume Diagram for 2200 rpm and 1300 rpm

Figure 7 excerpts, temperature attained by 1300rpm 1869.4K is slightly more than 2200rpm which is 1775.48K taking into consideration the greater amount of air forced into the chamber for 1300 rpm. Also, A/F ratio taken is approximately equal for both the cases should also be considered.

Figure 7: Temperature plot for 2200 rpm and 1300 rpm

We clearly see higher brake torque achieved by 1300rpm, which is an important requirement for heavy-duty applications. Also, we observe equal and symmetrical pattern for crake torque produced in both the cases throughout 720 Crank Angle Degree.

Figure 8: Brake Torque plot for 2200 rpm and 1300 rpm

Indicated Mean Effective Pressure suggests the output or work done by the engine. For both the cases, we see a sudden increase in IMEP during the expansion stroke, piston moving from TDC to BDC i.e. till Exhaust Valve Opening and then remains a sort of constant. A negative plot at 0 deg TDC, suggests the work is been done on the piston during the compression stroke.

Figure 9: IMEP for 2200 rpm and 1300 rpm

Nitric Oxide is a function of high temperature and intermediate equivalence ratio. Hence, we see a slightly more NOx to be formed in the case of 1300 rpm as the temperature produced is also high comparatively.

Figure 10: NO plot for 2200 rpm and 1300 rpm

Figure 11: Burned O2, CO2, H2O mass fraction plot for 2200 rpm and 1300 rpm

CONCLUSION

Through the CFD simulations, we were able to predict and achieve the given targets for Indian tractor application Engine well within 5%. Furthermore, the parametric study can be carried for the Engine with the same initial and boundary conditions for the optimization purpose.