Modeling of the Combustion Process for a Dual-fuel Diesel System

To reduce CO2 emissions, diesel engines operating in dual-fuel mode

use natural gas in the charge mixture to partly substitute the injected

diesel mass. For this mode of operation a phenomenological combustion

model based on extensive measurements in a rapid compression expan-

sion machine and a single-cylinder engine test bench was developed

and validated. The model reproduces the complex underlying processes

of dual-fuel diesel combustion and contributes to new knowledge of the

governing in-cylinder phenomena.

AUTHORS

Ömer Ünal, M. Sc. is Doctoral Candidate at

the Institute for Combus-tion Engines (IVK) of the

University of Stuttgart (Germany).

Dr.-Ing. Michael Grill is Head of 0-D/1-D Simu-lation at the Forschungs-

institut für Kraftfahrwesen und Fahrzeugmotoren

FKFS in Stuttgart (Germany).

Dr. sc. Sushant S. Pandurangi

is Postdoctoral Researcher at the Aerothermoche-

mistry and Combustion Systems Laboratory (LAV)

at ETH Zürich (Switzerland).

Dr. sc. techn. Yuri M. Wright

is Head of Engine CRFD at the Aerothermoche-

mistry and Combustion Systems Laboratory (LAV)

at ETH Zürich (Switzerland).

© University of Stuttgart

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1 MOTIVATION

Compressed Natural Gas (CNG) with methane as the main constit-uent is a viable alternative fuel to reduce CO2 emissions. Due to the favorable C/H atom ratio, methane produces approximately 20 % lower CO2 emissions for the same energy input compared to diesel. It is still possible to continue using the engine in normal diesel mode without restrictions. In dual-fuel mode, approximately 20 to 80 % of the diesel fuel is substituted by CNG, leading to very lean back-ground charge mixture (λ ≫ 2). This combustion mode is however distinct from gas engine dual-fuel operation, where small diesel pilot injections (1–2 % of the total energy input) are used merely to pro-duce ignition kernels and the combustion of the reactive charge occurs mainly through flame propagation. A dual-fuel diesel com-bustion model was developed in this research project. The model is based on comprehensive measurements from the research proj-ect [1] performed at the University of Stuttgart, as well as insights obtained from 3-D CFD calculations and tests on a Rapid Compres-sion Expansion Machine (RCEM) performed at ETH Zurich.

2 TEST BENCH AND MEASUREMENTS

Information about the test benches is given in TABLE 1. For phe-nomenological modelling of dual-fuel combustion, it is first nec-essary to identify the underlying combustion processes. To this end, the specific capabilities of each test bench were exploited: in the research engine, a substitution rate variation was performed. Here, the diesel amount injected is partly substituted by increas-ing amounts of methane in the charge gas, see Eq. 1):

Eq. 1 χ =

m ˙ Diesel ⋅ H u,Diesel _____________________ m ˙ CNG ⋅ H u,CNG + m ˙ Diesel ⋅ H u,Diesel

Meanwhile in the RCEM the parameters such as duration of injec-tion, ambient pressure, air–fuel ratio of the charge can be inde-pendently controlled (λCNG, see Eq. 2):

Eq. 2 λ CNG = m Luft ___________ m CNG ⋅ L min,CNG

FIGURE 1 (left) shows the overall dual-fuel Heat Release Rate (HRR) from the engine measurements, obtained using the mean pressure rate, for a variation in substitution rate. Under the typical conditions at Start of Injection (SoI) in diesel engines, presence of methane in the charge gas has no influence on the Ignition Delay (ID) and the initial combustion phase is similar to typical diesel operation, regardless of the substitution rate. In the later phases, increasing substitution of diesel by CNG results in significantly higher HRRs, caused by the entrainment of the reactive background mixture into the diesel spray. This highlights the importance of mod-elling the entrainment effects and the dependence on oxygen avail-ability. FIGURE 1 (right) shows similar trends in the RCEM, where

1 MOTIVATION

2 TEST BENCH AND MEASUREMENTS

3 MODELING

4 RESULTS AND VALIDATION

5 CONCLUSION

FIGURE 1 Heat release rate for a substitution rate variation in the engine (left) and independent variation of λCNG for different injection durations in the RCEM (right) (© University of Stuttgart | ETH Zürich)

Mixture Formation and Combustion

TABLE 1 Specifications of the research engine and the RCEM used (© University of Stuttgart)

Quantity Unit Engine RCEM

Engine name - MAn D0834 –

number of fired cylinders/valves

- 1/4 1

Stroke and bore mm 125 × 108 249 × 84

Cylinder volume cm³ 1145 1380

Compression ratio - 17.3 : 1 20 : 1

number of nozzle holes - 9 1

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four injection durations (300, 400, 700, 1000 µs energizing time) were used and λCNG was independently varied [2, 3]. As in the engine, at the typical diesel engine operating conditions (1000 K and 45 bar at SoI), no influence on the ID can be seen. This con-trasts strongly with findings reported at gas engine dual-fuel con-ditions, where SoI temperatures are typically below 850 K, where a significant influence of methane on the ID was found [4, 5].

Further insights into the governing processes of dual-fuel com-bustion can be obtained using propane variation, as shown by

results of a reaction kinetic calculation for a range of pressure and temperature. The addition of 10 % propane (by moles) does not significantly alter the laminar flame speed. While the laminar flame speeds are similar, the combustion duration is drasti-cally reduced in dual-fuel operation when propane is added, as seen in FIGURE 2. Combustion mechanisms based on a laminar- turbulent flame front alone are not sufficient to explain the short-ening of the combustion duration. Furthermore, when propane is added, high-frequency components in the pressure rate of

FIGURE 2 Heat release rate for a propane rate variation in the engine ( © University of Stuttgart)

FIGURE 3 Heat release rate for simultaneous variation of rail pressure and injection duration in the engine (left) and independent variation in the RCEM (right) for a diesel as well as dual-fuel case (λCNG = 2.1) (© University of Stuttgart | ETH Zürich)

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the individual-cycle cylinder pressures also increase, indicating self-ignition.

Diesel engine operation is characterized by significant spray-in-duced turbulence. Spray penetration length and mixture formation behavior are functions of rail pressure and injection duration. A longer penetration length leads to an enlarged spray plume surface, which should promote combustion of the background mix-ture by deflagrative mechanisms. If laminar-turbulent flame propa-gation occurs, it should therefore be influenced significantly by the rail pressure. As can be seen in FIGURE 3 (left) however, a rail pres-sure variation does not cause any remarkable differences in the HRR, and the initial combustion period and post-combustion are nearly identical. This suggests that other combustion mechanisms beyond flame propagation are present. Rail pressure was also var-ied in the RCEM, FIGURE 3 (right), where the duration of injection can be held constant. This causes, as expected, significantly higher HRRs until the end of injection, for diesel as well as dual-fuel operation. As in the engine, a sustained heat-release occurs here even long after the end of injection. As opposed to the engine, in the RCEM increas-ing rail pressure manifests itself in higher HRRs representing the higher momentum due to constant duration of injection and the con-sequently higher entrainment rate into the spray. Further, 3-D CFD simulations with a novel dual-fuel combustion model [6] have shown that the contribution of flame propagation under the conditions stud-ied is very low due to the typically fuel-lean background mixture, and that the combustion is instead dominated by large-scale fluid dynam-ics and turbulence effects induced by the spray.

The analysis of individual engine cycles showed an increase of the high-frequency components with increasing substitution rate. Applying a previously validated auto-ignition model for CNG mix-tures [7] indicates self-ignition at times when the combustion is accelerated. These observations strongly motivate the need for mod-elling self-ignition processes in the reactive background charge.

3 MODELING

Entrainment: Due to injection and mixture formation occurring simultaneously, the diesel and CNG combustion processes must compete with each other for oxygen. A λ model was developed from the diesel and oxygen on offer for this reason. To start with, care was taken that combustion with insufficient oxygen cannot occur. An increase in mixture richness was modeled between the moment of lack of for the CNG combustion and the moment of oxygen provision to λCNG (Eq. 2). Due to the disk approach used (according to [4]), each disk spans three different λDiesel regions. The volume of the disks in the outer diffusion region is calculated by the geometric spray cone angle and the diesel region with an air–fuel equivalence ratio of λDiesel > 1.1. This region is very lean and considering the diesel spray angle, λCNG and the spray model, large CNG mass is being calculated. However, the sudden com-bustion of the total mass of CNG in the lean region does not fit the concept. It can be assumed that there is a relationship between the quantity of diesel combusted in the lean region and the com-busted CNG. To verify this, the quantity of combusted CNG in this region is calculated, using the integral of the burnt diesel fuel quantity in relation to the rate (Eq. 3).

Eq. 3 r Diff,2 =

d m Diff,2 _________ c skal,E ⋅ m Diff,2

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As soon as this value exceeds a limit cGrenz, λCNG takes on the form of Eq. 4.

Eq. 4 λ DF = λ CNG ⋅ r Diff,2

The entrainment combustion within the diesel spray can be cal-culated from the λ model by using Eq. 5.

Eq. 5

d Q b,Ent _____ dϕ = λ DF ⋅ L min ⋅ d m b,Diesel ______ dϕ ⋅

m CNG ____ m total ⋅ H U,CNG

Flame propagation: For a number of operating points, the pres-ence of a slow laminar-turbulent flame propagation can be assumed. However, many operating points switch quickly into the mode of accelerated combustion due to self-ignition in the CNG background mixture. For this reason, detailed modelling of the flame surface was not undertaken. The flame propagation was modeled at this point using a simplified assumption of hemispher-ical flame propagation (Eq. 6, Eq. 7).

Eq. 6 d m e ___ dϕ = ρ UB ⋅ A F ⋅ C skal,def ⋅ u e

Eq. 7 d Q b,Deflag.  _______ dϕ =

∫ ( d m e − d m B ) _________ τ ⋅ dτ ___ dϕ ⋅ H u,CNG

Here the flame surface was calculated using the number of injec-tion spray holes, which incorporates a consideration of the cyl-inder shape. The time of flame propagation was set as the end of injection. Once a size which overlaps two flames surfaces is reached, the surface reduction is calculated using a scaling

parameter. The deflagrative components of dual fuel combustion result from the laws of spark ignition flame propagation.

Volumetric combustion: The combustion of the CNG mass by self-ignition is governed by the ignition integral from [7]. As soon as the average tem perature in the unburnt mixture exceeds the self-ignition limit of the CNG, the CNG burns with volumetric com-bustion. Here, the total mass fraction of combustible CNG is tab-ulated sigmoidally. To take the lack of temperature homogeneity in the combustion chamber into account, a temperature distribution around the average unburnt temperature is calculated for each moment (Eq. 8).

Eq. 8 T σ = T UB + σ

If self-ignition occurs for a certain temperature history, the appropri-ate mass of CNG is burned. The combustion via self-ignition is cal-culated from the change in burned mass rate for self-ignition (Eq. 9).

Eq. 9 d Q b,Volume ________ dϕ =

Δ Sigmoid ⋅ m CNG __________ dϕ ⋅ H u,CNG

4 RESULTS AND VALIDATION

FIGURE 4 shows the dual-fuel combustion process including the breakdown of single combustion mechanisms and the rate of die-sel injection. This operating point has a substitution rate of 50 % at a medium speed and high part-load. The air ratio in the back-ground mixture λCNG is 2.86 (pSoI / TSoI [bar/ K] 88 / 1067) and accordingly the proportion of the deflagrative combustion compo-nent is low. It can be seen that at relatively high diesel levels, the combustion of CNG within the spray envelope takes the largest amount and the oxygen-limited initial phase is well reproduced.

FIGURE 4 Heat release rate of modeled dual-fuel combustion with single com-bustion mechanisms, case 224-05 (© University of Stuttgart)

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The volumetric combustion starts during the injection phase and converts an amount by self-ignition.

FIGURE 5 This operating point has a 60 % substitution rate at medium speed and high part-load. The air ratio in the background mixture λCNG is 3.1 (pSoI / TSoI [bar/K] 102.4 / 1021.7). According to the lean background mixture, the deflagrative component is also low at this operating point. The self-ignition model calculates in this case the reaching of the self-ignition criterion in the falling side of the entrainment combustion and thus results in a rising side, which takes place by volumetric combustion.

5 CONCLUSION

This project investigates dual-fuel combustion in diesel engines, where energy substitution is carried out by premixing methane into the background air. It is seen that the ignition delay of the diesel spray is unaffected by the methane content for typical conditions in commercial diesel engines, in strong contrast to gas engine dual-fuel conditions. The first phase of dual-fuel combustion is determined by a lack of oxygen. Methane is strongly entrained into the diesel envelope and enhances the heat release rate at later stages. Flame propagation in the background mixture is often very low due to the very lean gas mixtures (λ ≫ 2), and spray-induced fluid mechanic effects therefore play an important role. High-fre-quency oscillations typical of engine “knocking” are seen on the individual-cycle pressure traces, indicating self-ignition occurring in the background mixture, also supported by reaction kinetic cal-culations and by the addition of propane. All processes have been accounted for in the developed phenomenological model showing good agreement with the experimental data.

REFERENCES[1] Tänzler, A. G.: Experimentelle Untersuchung eines Dual-Fuel-brennverfah-rens für schwere nutzfahrzeugmotoren. Wiesbaden: Springer Fachmedien, 2017 [2] Srna, A.: Experimental Characterisation of Pilot-Fuel Ignition, Combustion, and Soot Formation in Dual-Fuel Combustion Systems. Zürich, ETH, PhD thesis, 2018

[3] Srna, A.; bruneaux, G.¸ von Rotz, b.¸ bombach, R.; Herrmann, K.; bou-louchos, K.: Optical Investigation of Sooting Propensity of n-Dodecane Pilot/Lean-Premixed Methane Dual-Fuel Combustion in a Rapid Compression Expan-sion Machine. SAE Technical Paper 2018-01-0258, 2018[4] Schlatter, S.; Schneider, b.; Wright, Y. M.; boulouchos, K.: n-heptane micro pilot assisted methane combustion in a Rapid Compression Expansion Machine. In: Fuel, 179 (2016), pp. 339–352[5] Srna, A.; bolla, M.; Wright, Y. M.; bombach, R.; Herrmann, K.; Pandurangi, S. S.; et al.: Effect of Methane on Pilot-Fuel Auto-Ignition in Dual-Fuel Engines. Proceedings of the Combustion Institute 37 (2019), no. 4, pp. 4741-4749[6] Seddik, O.; Pandurangi, S. S.; Wright, Y. M.; bolla, M.; boulouchos, K.; Srna, A.: Flamelet Generated Manifolds applied to dual-fuel combustion of lean meth-ane/air mixtures at engine relevant conditions ignited by n-dodecane micro pilot sprays. SAE Technical Paper 2019-01-1163, 2019[7] Urban, L.; et al.: Simulation of Autoignition, Knock and Combustion for Methane- based Fuel., SAE Technical Paper 2017-01-2186, 2017[8] Rether, D; Grill, M.; Schmid, A.; bargende, M.: Quasi-Dimensional Modeling of CI-Combustion with Multiple Pilot- and Post Injections. In: SAE International Journal of Engines 3 (2010), no. 1, pp. 12-27

THANKSThis article is based on research project 1219 “Dual Fuel CI Engines” undertaken by the FVV (Research Association for Combustion Engines e. V.) and performed by the Institute for Combustion Engines (IVK) at Universität Stuttgart led by Prof. Dr.-Ing. Michael Bargende and by the Aerothermochemistry and Combustion Systems Laboratory (LAV) at ETH Zurich led by Prof. Dr. Konstantinos Boulouchos. We would like to thank the project chairman Dr. Christian Barba (Robert Bosch GmbH) and all members of the user committee for their support. The project was carried out in the framework of the industrial collective research program (IGF) (No. 18983N). It was supported by the Federal Ministry for Economic Affairs and Energy (BMWi) through the German Federation of Industrial Research Associations e. V. (AiF) based on a decision taken by the German Bundestag. The project was funded in part by the FVV from own resources (No. 6012192) and the Swiss Federal Office of Energy (SFOE) (No. SI/500970-01). The authors gratefully acknowledge the support received. In addition to the listed authors, this article is co-authored by Dr. Aleš Srna, M.Sc. Omar Seddik, Prof. Dr. Konstantinos Boulouchos and Prof. Dr.-Ing. Michael Bargende.

FIGURE 5 Heat release rate of modeled Dual- Fuel combustion with single combustion mecha-nisms, case 224-04 (© University of Stuttgart)

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