Printer Friendly

Development of multi cylinder turbocharged natural gas engine for heavy duty application.


CNG has recently seen increased penetration within the automotive industry. Due to recent sanctions on diesel fuelled vehicles, manufactures have again shifted their attention to natural gas as a suitable alternative. Turbocharging of SI engines has seen widespread application due to its benefit in terms of engine downsizing and increasing engine performance [1].

This paper discusses the methodology involved in development of a multi cylinder turbocharged natural gas engine from an existing diesel engine. Various parameters such as valve timing, intake volume, runner length, etc. were studied using 1D simulation tool GT power and based on their results an optimized configuration was selected and a proto engine was built. Electronic throttle body was used to give better transient performance and emission control. Turbocharger selection and its location plays a critical role. Turbocharger Wastegate actuator trials were conducted to select optimum actuator to restrict boost enough to meet power target. Various engine related MAP like main lambda, volumetric efficiency, MBT timing, Primary ignition timing, etc. were calibrated and saved in EMS. The full load curve was calibrated using EMS to meet performance target of 85 kW rated power and 285 Nm max torque. Further, through precise calibration of various ECU tables the engine managed to meet BS IV emission norms.

CITATION: Thipse, S., Dsouza, A., Sonawane, S., Rairikar, S. et al., "Development of Multi Cylinder Turbocharged Natural Gas Engine for Heavy Duty Application," SAE Int. J. Engines 10(1):2017


Ever-increasing cost & limited resources of petroleum based fuels and the stringent emission norm has put lot of pressure on the automobiles manufacturer and researchers to improve the fuel economy and to reduce the exhaust gas emission of their vehicles [2]. This has lead manufacturers and scientist to experiment with newer technologies. Downsizing of engines have seen tremendous scope when it comes to SI fuelled engines. Use of alternative fuels along with downsizing of engines has helped engineer's meets stringent emission norms while maintaining vehicle performance and also improving fuel efficiency [3, 4].

CNG due to its low C/H ratio is known for emitting lower amounts of C[O.sub.2] than gasoline [3]. The improvement of fuel efficiency will directly translate into lesser C[O.sub.2] emission which will give manufacturers an edge once C[O.sub.2] is included as a regulated pollutant in the forthcoming BSVI emission norms that is to be implemented by year 2020 in India.

Turbocharging an engine has long been considered the best way to improve engine performance. Its application in diesel engines is seasoned but has rarely been used in SI engines. The presence of a throttle body in the intake stream of an SI engine can cause functional problems for a turbocharger which were investigated and solutions to avoid the same were implemented. The constant opening and closing of the throttle causes continuous zones of low and very high bottled up pressure before and after the compressor. Not venting the high pressure after the compressor when the throttle body is closed can lead to damage of the compressor bearings. CNG being a gaseous fuel of higher calorific value sees higher exhaust temperatures. Diesel engine turbocharger are designed to withstand exhaust temperatures of up to 705[degrees]C. Turbocharging a CNG engine can see exhaust temperatures exceed 800[degrees]C.

The base engine was a multi cylinder turbocharged diesel engine used on heavy duty above 3.5T applications. The base engine was converted to a turbocharged dedicated CNG engine for the same application. This paper discusses the different phases, calibration activities and experimentation that was involved in the development.


Base diesel and proposed CNG engine specifications are mentioned in the table 1. It can be seen that while converting the base diesel engine

to CNG application the rated power target was increased by around 20% and there was change in the required torque target too. To achieve the desired target the base BMEP level was increased from 9.9 bar to 11 bar. Hence it was indeed required to check the suitability of existing base diesel engine's components for CNG operation.


The new engine is developed to meet BSIV engine emission norms. And those are as per Table-2. Applicable cycle is ETC when checked for normal engine testing with no DTC/fault in the system or component.


Intake Manifold and Cylinder Head

Initially the existing cylinder head and intake manifold was tested for swirl. This Intake manifold has small volume and cylinder head port was designed to create swirl which is very critical for diesel engines. CNG engines requires intake manifold with larger volume for smooth operation and greater torque response [1, 5]. Below figure shows that the mean swirl number increases with the use of short runner intake manifold i.e. diesel engine intake manifold. This higher swirl is not desirable for CNG application.

to overcome this issue, the new intake manifold was proposed on the basis of 1D and 3D simulation which has bigger plenum volume and runner length compared to diesel intake manifold which will help in storing higher charge/ air as a reservoir. Also the runner length and its orientation is finalized in such a way that it kills the swirl and improve the overall volumetric efficiency. To accommodate spark in place of diesel injector a small protrusion is created in intake port area which in turn helps to reduce the swirl. Figure 2 shows the mean swirl number comparison with new and existing intake manifold.

Cooling and Lubrication System

The existing water pump was designed to meet the cooling requirement for the 70kW engine. The cooling requirement for the proposed engine was calculated and it was found that water flow requirement is 211 LPM for rated Power & 115 LPM for Max torque. So the existing water pump was not suitable for the CNG application. So accordingly new water pump is designed and used further.

The same case was with the lubrication system. The existing oil pump was also not capable for providing required oil flow with the desired pressure. So accordingly new requirement was given and new oil pump was developed. The existing oil was also reviewed. Oil with low ash additives is suggested which will give protection against carbon deposit. This oil caters low sulphur to protect after-treatment device. Also this offer outstanding resistance to pre-ignition and protection against carbon deposit on valves and spark plugs.

Piston & Ring Pack

Piston and ring packs are one of the major contributing components to achieve desired performance and emission. Piston & ring pack are reviewed for the following point of views:

* minimize dead volumes

* minimize friction losses

* minimize oil consumption

* improve durability

Considering the increased BMEP of developed CNG engine it is recommended to use symmetric / asymmetric barrel type compression ring. The key stone ring used in diesel engine is not recommended considering its scraping operation due to low clearance & high temperature generated in CNG engines.

Compression Ratio

For existing diesel engine CR is 18:1. For converting the same engine to run for CNG, Compression Ratio needs to be reduced. For proposed CNG engine two different CR ratios (10.5 & 11) were initially proposed. However to achieve the more power in CNG engine, boosting in the form of turbocharger is considered. Accordingly air requirement is calculated & it is observed that the inlet air temperature is the key factor to decide the CR to avoid auto ignition of CNG fuel [5]. While considering the worst case situation, atmospheric intake air temperature is ranging from 45[degrees]C to 50[degrees]C & boost pressure of around 1.5 bars, there are chances of significant increase in compression temperature. So in the present engine configuration Compression Ratio (CR) 10.5: 1 was used for further optimization. And if required CR of 11:1 will be considered.

Combustion Chamber

The existing shape of the combustion chamber (Piston Bowl) in the diesel engine is re-entrant type, which is not suitable for CNG application. Combustion chamber for CNG engine is designed based on parameters like turbulence, low swirl and squish [5, 6]. For natural gas due to high activation energy, the laminar flame speed is low which results in longer combustion duration. Thus the total combustion period becomes prolonged as compared to diesel and petrol. This leads to loss in efficiency. Longer the total combustion time, higher chances of the remaining unburned mixture to undergo pre-flame reactions and self-ignite.

Fuel burning before top dead center increases the work required for compression, while that of burning late in the cycle performs less work on the piston during expansion. This problem of low flame speed can be solved by two ways. The first is that the ignition timing could be advanced. The other option is the development of combustion chamber specifically for natural gas operation by Increasing turbulence & squish in the combustion chamber to increase flame speed. A squish generated charge motion combustion chamber had its effects on the burning rates. High levels of turbulence generated from the squish affected the faster burning rates, which resulted in improvement in thermal efficiencies. Figure 3 shows the typical combustion bowl shape for diesel and CNG engine.

Piston Pin

Due to change in the peak firing pressure of the engine, it is essential to check the strength of the piston pin. FE analysis is done for checking structural integrity of the piston pin.

The analysis is based on the simple assumption that; pin will act as a simply supported beam when assembled in the piston. Thus, pin is held on its bottom side in patch of 120 degrees. (Assuming that to be the contact patch.) Gas load is applied on the pin, at each node. This force is also applied on contact patch of 120 degrees. Appropriate material properties are assigned to the material for determining the stress and the displacements. Analysis is done on three different pressure conditions (90bar, 100bar, 110bar). Figure 4. Shows the case set up for the piston pin analysis.

The existing diesel piston pin is having 5.5 mm wall thickness and 74mm length. The max. Stress observed at 110 bar combustion pressure is 454 MPa which is greater than maximum fatigue strength of the material selected i.e. 420 MPa and hence existing piston pin was not suitable for the CNG engine application.

So accordingly the suggestions were made and the Pin thickness was increased to 7mm and length of the pin was increased by 3mm and again the simulation was done. It was found that max. Stress observed at 110 bar combustion pressure is 340 MPa, which was very much below the acceptable limit and hence it was accepted.

Table 3. Shows the difference between the maximum stress level and deflection for existing and modified piston pin.

Stress due to boundaries is not the realistic stress and hence can be neglected

Catalytic Convertor

The base engine was not having any after treatment system as the applicable norms was BSIII. For this proposed BSIV CNG engine the three way catalyst was recommended. The light off temperature suggested by the catalytic converter manufacturer was around 350[degrees]C. The loading of the new catalytic converter was 60 gm/[ft.sup.3] and the relative concentration for Pt/Pd/Rh was 0/90/10 while keeping the cell density constant at 400 cpsi.


As CNG engine is having more power target than the base diesel engine, turbocharger is the prime need to fulfill the power target. The proposed CNG engine is running on stoichiometric principle so for calculating the required air flow rate initially the SFC target was set for each engine operating speed. Then on the basis of fuel consumption the required air flow was determined. For this engine intercooler was also considered as after turbocharging the temperature of the intake air will rise drastically. This rise in temperature may further reduce the air flow and may cause increase in the compression temperature which may then lead to knocking. So the air temperature after intercooler was fixed and further calculations for turbocharger was done.

Initially two different calculations one without EGR and other with EGR were done. The actual test results of the selected turbocharger are discussed in the subsequent part of this paper.


1D simulation model is built in the GT power software. This model is built to predict the effect of different valve timing, plenum volume and runner length of Intake Manifold on engine performance [7]. A schematic engine model is shown in Figure.7 Actual engine components are schematically represented by symbols. Geometrical and operating boundary conditions are applied as an input to the model. Simulation runs are taken at full throttle condition for maximum power engine speed, maximum torque engine speeds and idle speed.

Valve Timing Study

Existing valve timing was taken as a base for further investigation. To have the minimum changes in cam shaft profile, only cam lobe shifting was considered & accordingly intake & exhaust valve timing was retarded & advanced by 10[degrees]. So there are total 9 cases was studied with different IVO, IVC, EVO, EVC events. The effect of various valve timings on the volumetric efficiency, BSFC, Residual fraction, Air flow rate, and intake pressure is shown in figure no 8 to 13 respectively.

It can be seen from the figures that advancing the exhaust timing improves the BSFC but penalty in terms of lower volumetric efficiency noticed. Retarding the exhaust timing improves the volumetric efficiency but penalty in BSFC observed. Retarding Exhaust timing and Keeping Intake timings the same seen to give best volumetric efficiency results and 2% drop in BSFC in comparison to best result. Base Valve timing seen to give optimum volumetric efficiency and BSFC results. So considering the all advantages and disadvantages base timings and option 4 in which exhaust valve timing 10[degrees] retard was found suitable.


Effect of Plenum Volume

In this study the plenum volume is varied keeping the runner cross section and length constant. Three different volumes have been studied and its effect on volumetric efficiency and engine performance are plotted. The three different manifold volumes considered are 100%, 70% and 130% of designed volume [7]. Comparison of effect of various plenum volume on engine performance for is shown in figure no 13 to 15.

It is seen that there is minimal improvement observed with increasing Plenum volume. So it is not advisable to increase plenum volume to get minimal improvement of volumetric efficiency. Whereas there is drop in volumetric efficiency and deterioration in BSFC observed due to reduced plenum volume.

Effect of Runner Length

Three different runner lengths are studied here. The base length is taken as L. Three different cases L, L+100mm and L-100 mm on the performance and volumetric efficiency is studied [6]. The graphs are plotted as follows. Comparison of effect of various runner length on engine performance for is shown in figure no 16 to 18.

It is observed that there is minimal improvement (Maximum 2%) observed with increasing runner length. So it is not advisable to increase runner length by 100mm to get only 2% of volumetric efficiency. As the height of the intake manifold with 300 mm runner length is too big to accommodate it on the engine due to packaging constraints. Whereas there is drop in volumetric efficiency and air flow with 100 mm runner length. Also CNG injector positioning on shorter runner length is having engine boundary limitations so it is not suitable. Hence runner length (190mm) was found suitable for this engine operation.


Experimental visualization of air flow patterns due to Intake manifold wall is very difficult. Computational fluid dynamic (CFD) techniques are used for determining the flow though pipes. The flow through each runner is found out. The Intake Manifold which found suitable in 1D simulation analysis is further taken up for 3D simulation work. In flow analysis the flow though each runner was determined and it was found out that Maximum flow variation was found around 7% compared to fifth runner. And overall runner to runner variation was around 5% which shown in figure 19.


After reviewing the major engine components/parameters now the time was to build an actual engine and calibrate it for BSIV emission norms on ETC. Schematic layout for proposed CNG engine is shown in figure 21. Throttle body opening signal will be provided to the dump valve and as per the boost requirement, bypass valve will be opened and compressed air will be bypassed before throttle body to the compressor side of turbocharger. Dump valve is introduced into the system to safeguard the throttle body and compressor of the turbocharger.


The components were developed as per the design guidelines and engine was assembled and tested on eddy current and transient dynamometer. All the major calibration activities were completed on eddy current dynamometer. Below are the activities performed while engine calibration phase.

Wastegate Actuator Selection

Depending on the turbocharger calculations the required boost to achieve desired engine performance was initially informed to turbocharger manufacturer. But because of higher exhaust temperature in the range of 800-850[degrees]C, supplier was having only one set of compressor and turbine housing. So it was decided to select the actuator in such a way that it will operate the Wastegate to get the desired performance. Hence Wastegate actuators vis-a-vis 120N/mm, 100N/mm, 80N/mm, 65N/mm and 55N/mm were made available for the boost study. Different actuators with various stiffness are shown in below figure 22.

Throttle body with 52mm diameter was considered for initial trials of Wastegate actuators. Peak Boost pressure of 2.4, 2.2 and 2.0 bar was observed with 120N/mm, 100N/mm, and 80N/mm actuators respectively with very minimum throttle body opening and also exhaust temperatures was also above 850[degrees]C.

Boost pressure of 1.5 observed with 65 N/mm WG actuator while with 55 N/mm actuator a peak boost pressure of 1.44 Bar was achieved. Throttle opening increased to 39% with 65 N/mm WG actuator and further increased to 44% with 55 N/mm WG actuator. Maximum exhaust temperature was 823[degrees]C with 65 N/mm WG actuator while it decreased to 816[degrees]C with 55 N/mm WG actuator at rated RPM with back pressure control.

So it was decided to go with either 55N/mm or 65N/mm actuators.

Engine performance with 55N/mm or 65N/mm waste gate actuators and dia. 52mm throttle body is shown in figure 23 and 24 respectively.

Electronic Throttle Body Selection

As seen from the FTP trials maximum 44% of throttle opening was enough to achieve the desired output. That means the throttle body was too sensitive for practical work. So it was decided to check the engine performance with 44mm dia. electronic throttle body. It was observed that Wastegate actuator with 55N/mm stiffness and 44mm dia. throttle body gives the desired power performance. And 100% throttle opening was required to get the desired performance. Engine performance with 55N/mm waste gate actuators and dia. 44 mm throttle body is shown in figure 25

Fuel Trim

The purpose of the fuel trim is to provide the catalytic converter with a rich and lean air fuel mixture. A rich mixture (lack of oxygen) is needed by the catalytic converter to reduce NOX exhaust emissions. The catalytic converter must also receive a lean mixture (excessive of oxygen) to oxidize the HC and CO into harmless C[O.sub.2] and [H.sub.2]O. If the exhaust is always rich, the cat con cannot reduce the CO and HC emissions. If exhaust is always lean the cat con cannot reduce N[O.sub.X] emissions therefore air fuel ratio must alternate between rich and lean. The computer is designed to provide this alternating mixture by using the oxygen sensor, short term and long term fuel trim program to accomplish this feat.

Volumetric Efficiency Mapping

Volumetric efficiency is described as the ratio of the volume of fluid actually displaced by a piston or plunger to its swept volume. It defines the breathing capacity of the engine. CNG engines in comparison to its gasoline counterparts suffer from lower volumetric efficiency due to displacement of air by natural gas. By proper location of the fuel injectors this loss can be minimized.

The Volumetric Efficiency map in the ECU is optimized in such a way that correct amount of fuel is metered at all instances so as to maintain the correct lambda. The fuel metering is controlled by the lambda sensors in the exhaust system.

During optimization the volumetric efficiency map was calibrated in such a way that the correction that the ECU would have to provide to meter accurate amount of fuel would be minimal. The three dimensional plot of the calibrated volumetric efficiency is shown in figure 26 .The volumetric efficiency is mapped with respect to engine speed and manifold absolute pressure (MAP)

Ignition Timing mapping

Ignition timing affects the overall performance of an engine. Too advance a timing leads to higher NOx emissions while too retard a timing causes a drop in power and torque. The timing should be optimized such that engine performance and emissions are not compromised.

Ignition timing calibration consists of optimizing two ECU maps.

* Maximum Brake Torque (MBT) Ignition Timing

* Primary Ignition timing

Maximum Brake Torque (MBT) timing is the use of optimal ignition timing to take advantage of an internal combustion engines maximum power and efficiency. This timing is used during WOT conditions so that engine delivers maximum power and torque. The MBT timing was calibrated for maximum torque. The Torque and exhaust temperatures were noted down and used to populate the MBT Torque table and Base exhaust temperature table. The three dimensional plot of the calibrated MBT timing, MBT torque and Base Exhaust temperature are as shown in figures 27, 28 and 29 respectively. All calibration tables are mapped with respect to engine speed and engine load.

Primary Ignition timing table was calibrated to optimize the part load zones and also reduce emission for the transient cycle. The three dimensional plot of the calibrated Primary ignition timing, is shown in figure 30.

This table is mapped with respect to engine speed and engine load

Drag Torque Calibration

Drag Torque table calibration is used to define the engine frictional torque and pumping losses under warmed-up conditions. It is used to calculate the proper throttle opening during deceleration fuel cut-off conditions, and also as part of the engine torque estimate calculation based on efficiency. This table must be mapped using a motoring dynamometer. The engine is operated at the desired speed and load location with fuel, then the fuel is cut and the torque required to motor the engine is measured. The three dimensional plot of the calibrated Primary ignition timing, is shown in figure 31

The table is mapped with respect to engine speed and Engine Coolant Temp (ECT)

Commanded Lambda Calibration

Main Lambda table was calibrated for minimum emissions and best SFC at part load points and to reduce maximum exhaust temperature at full load rated points. The three dimensional plot of the calibrated Main Lambda, is shown in figure 32

The table is mapped with respect to engine speed and Rel. Torque.

Accelerator Paddle Map Calibration

This Map was calibrated to give best transient response on the ETC cycle. The three dimensional plot of the calibrated accelerator Paddle Map, is as shown in figure 33

The table is mapped with respect to engine speed and accelerator Paddle position.

Dwell Timing Calibration

Dwell is the amount of time required to charge an inductive coil to its maximum energy level. In terms of modern engine control, dwell is defined in milliseconds. Electronic ignition is much better at regulating dwell than distributor based engines this calibration process will require a scope-meter with current probe and a laptop to communicate with ECU. Using a breakout box will allow easy access to probe wire of the ECM pin-out. Measure the operating dwell time and operating supply voltage to achieve the operating current using the test setup at various calibration dwell time. Using this setup the nominal dwell timing table was calibrated and shown in figure 34.

Desired Torque Calibration

This table was calibrated to attain maximum torque curve as desired by the manufacturer. The plot of the calibrated desired torque, is as follows: The table are mapped with respect to engine speed and desired torque.

Max Throttle Position Calibration

This table is calibrated for throttle position at which maximum torque is observed. The plot of the calibrated Max throttle position, is as follows: The table are mapped with respect to engine speed and Throttle %.

Engine Full Throttle Performance Test

After all the calibration work the engine is tested for the full throttle performnce once again. The results for enigne performance are as shown in figures 38 to 40.

The experimental setup consists of the multi-cylinder engine coupled to a dynamometer. The engine and dynamometer are mounted on the test rig. The test cell is equipped with conditioning air system needed as per the testing standards for carrying out the testing as per standard atmospheric conditions. The engine test set up on is shown in figure 37.

Engine Emission Test

After the exhaustive calibration work the engine was tested on transient dynamometer on ETC. The emission results were much below the engineering target. After the initial emission trails the engine was again fine-tuned during the OBD calibration work. And after the OBD-II trials the emission level was retained well within the limit.


The following conclusions may be drawn from the present exercise

1. While converting the base diesel engine for CNG operation various components/parameters like Intake manifold design, Exhaust manifold design, Water pump and Oil pump capacity, Compression ratio, Piston and ring pack design, Piston pin design, Oil grade, turbocharger, Valve timings, needs to be reviewed.

2. Simulation software helps us in building the existing engine model and by varying the required parameters we can easily get its effect. This approach helps us in deciding the optimum design rather than actual validation on engine. So overall time and project cost can be reduced.

3. Turbocharger selection is one of the critical activity, because it is seen that by changing the value of actuator stiffness the whole performance curves get shifted.

4. The constant opening and closing of the throttle causes continuous zones of low and very high bottled up pressure before and after the compressor. Not venting the high pressure after the compressor when the throttle body is closed can lead to damage of the compressor bearings. So provision of dump valve is mandatory where both turbocharger and throttle body is present.

5. ECU control MAPs like Volumetric Efficiency, MBT Timing, Primary Ignition timing, Torque table, Drag torque, Port Air temperature model, Exhaust temperature model, Injection timing, etc. needs to be calibrate for better transient response and less emissions.

6. The developed CNG engine from existing diesel engine met the rated power and rated torque targets with a flat torque range over the engine speed range.

7. Closed loop control of CNG engines along with the three-way catalyst results in a large reduction in emissions of NOx, CO and HC. The engine meets BS IV emission regulations with significant margin easily.

8. CNG being a gaseous fuel of higher calorific value sees higher exhaust temperatures. In this engine the maximum exhaust temperature was around 800[degrees]C. But by richening the air fuel mixture at the identified zone the exhaust temperature can be reduced.


[1.] Midhun, V., Karthikeyan, S., Nagarajan, ., Krishnan, S. et al., "Development of CNG Injection Engine to Meet Future Euro-V Emission Norms for LCV Applications," SAE Technical Paper 2011-26-0002, 2011, doi:10.4271/2011-26-0002.

[2.] Khatri, D. and Rungta, P., "Development and Evaluation of a Multipoint Gas Injection System for a Passenger Car," SAE Technical Paper 2008-28-0067, 2008, doi:10.4271/2008-28-0067.

[3.] Semin et. al., "A technical review of compressed natural gas as an alternative fuel for internal combustion engines", American Journal of Engineering And applied science, (2008) 302-311

[4.] Iyengar, K., Kulkarni, D., Chopra, N., Palkar, A. et al., "Development of BS-III CNG Engine for a Light Commercial Vehicle," SAE Technical Paper 2007-26-028, 2007, doi:10.4271/2007-26-028.

[5.] Mohite, J., Suple, P., Wanpal, A., Chougule, N. et al., "Development of BS-III CNG Engine for Heavy Commercial Vehicle," SAE Technical Paper 2009-26-0038, 2009, doi:10.4271/2009-26-0038.

[6.] Midhun, V., Karthikeyan, S., Krishnan, S., Tyagi, A. et al., "Development of Three Cylinder CNG Engine for LCV Application," SAE Technical Paper 2013-26-0009, 2013, doi:10.4271/2013-26-0009.

[7.] Pai, D., Singh, H., and Muhammed, P., "Simulation Based Approach for Optimization of Intake Manifold," SAE Technical Paper 2011-26-0074, 2011, doi:10.4271/2011-26-0074.


Dr. S. S. Thipse

Deputy Director

Powertrain Engineering Division

The Automotive Research Association of India (ARAI)

S. No. 102, Vetal Hill, Off Paud road, Kothrud Pune-38 Maharashtra, India 091-020-30231434


Mr. Kamlesh Bhandari

Divisional Manager

Engine Design, Development and Testing

Force Motors Ltd

Mumbai - Pune Road Akurdi, Pune 411 035 Maharashtra, India

Tel : +91 20 2747 4613

Mob: +91 8805017145


The authors would like to thank Smt. R.H. Urdhwareshe, Director of ARAI, Mr. S. S. Ramdasi and his design team, Mr. N. H. Walke and his simulation team, and all colleagues at PTE division for their support and encouragement during the project. We are also thankful to M/s FML, India for providing opportunity to work for them on this project. We are also thankful to Mr. Vishal Singhal and Mr. Prashant Pawar of M/S Advantek for their extended support during the engine calibration work.


BMEP - Brake Mean Effective Pressure

BSFC - Brake Specific Fuel Consumption

CFD - Computational Fluid Dynamics

CNG - Compressed Natural Gas

CR - Compression ratio

DTC - Diagnostic Trouble Code

ECM - Engine Control Module

ECU - Engine Control Unit

EMS - Engine Management System

ETC - European Transient Cycle

FTP - Full Throttle Performance

GT - Gamma Technology

LPM - Litres Per Minute.

MBT - Maximum Brake Torque

SI - Spark Ignition

WOT - Wide Open Throttle

Sukrut S Thipse, Ashwin Dsouza, Shailesh B Sonawane, S D Rairikar, Kishorkumar Kavathekar, and Neelkanth Marathe

Automotive Research Association of India

Balasaheb Shinde, Sudhindra Kadkol, Kamlesh Bhandari, and Mandar Joshi

Force Motors, Ltd.

Table 1. Engine Specification

                   Base Diesel Engine  Proposed CNG Engine

No of Cylinder     5                   5
Cubic Capacity     3.25 Ltr            3.25 Ltr
Power (kW @ RPM)   70 @ 3000           85 @ 3000 (3400 initially)
Torque (Nm @ RPM)  255 @ 1400-2000     284 @ 1400-2000
Injection System   VE Mechanical       Multi Point Fuel Injection

Table 2. Emission norms

             BSIV Emission Norms

Monitored    CO                          NMHC
             (g/kWh)                     (g/kWh)

Legislation                         4                           0.55
Limit (LL)
Engineering  [less than or equal to]3.2  [less than or equal to]0.44

Monitored    CH4                          NOx
             (g/kWh)                      (g/kWh)

Legislation                         1.1                          3.5
Limit (LL)
Engineering  [less than or equal to]0.88  [less than or equal to]2.8

Monitored    PM

Legislation  Not
Limit (LL)   Applicable
Engineering  Not
Target       Applicable

Table 3. Piston Pin Simulation Result Summary

                Deformation at the    Max. Stress (MPa)
                loading boundary
Case  Analysis  (Existing  (Modified  (Existing  (Modified
      Pressure  Pin)       Pin)       Pin)       Pin)

1      90       0.0211     0.0134     362        278
2     100       0.0223     0.0149     402        309.1
3     110       0.0251     0.0164     454        340

                Stress Due to
                Boundaries (MPa)

Case  Analysis  (Existing  (Modified
      Pressure  Pin)       Pin)

1      90       792        712
2     100       874        791
3     110       900        870
COPYRIGHT 2017 SAE International
No portion of this article can be reproduced without the express written permission from the copyright holder.
Copyright 2017 Gale, Cengage Learning. All rights reserved.

Article Details
Printer friendly Cite/link Email Feedback
Author:Thipse, Sukrut S.; Dsouza, Ashwin; Sonawane, Shailesh B.; Rairikar, S.D.; Kavathekar, Kishorkumar; M
Publication:SAE International Journal of Engines
Article Type:Report
Date:Feb 1, 2017
Previous Article:Achieving Bharat Stage VI emissions regulations while improving fuel economy with the Opposed-Piston Engine.
Next Article:Design and analysis of a fuel preheating device for evaluation of ethanol based biofuel blends in a diesel engine application.

Terms of use | Privacy policy | Copyright © 2020 Farlex, Inc. | Feedback | For webmasters