Effect of turbo charging and steam injection methods on the performance of a Miller cycle diesel engine (MCDE)

Accepted Manuscript Research Paper Effect of Turbo Charging and Steam Injection Methods on the Performance of a Miller Cycle Diesel Engine (MCDE) Guven Gonca, Bahri Sahin PII: DOI: Reference:

S1359-4311(17)30879-7 http://dx.doi.org/10.1016/j.applthermaleng.2017.02.039 ATE 9920

To appear in:

Applied Thermal Engineering

Received Date: Revised Date: Accepted Date:

17 June 2016 31 January 2017 9 February 2017

Please cite this article as: G. Gonca, B. Sahin, Effect of Turbo Charging and Steam Injection Methods on the Performance of a Miller Cycle Diesel Engine (MCDE), Applied Thermal Engineering (2017), doi: http://dx.doi.org/ 10.1016/j.applthermaleng.2017.02.039

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Effect of Turbo Charging and Steam Injection Methods on the Performance of a Miller Cycle Diesel Engine (MCDE) Guven Goncaa1*, Bahri Sahina, a

Yildiz Technical University, Naval Arch. and Marine Eng. Depart, Besiktas, Istanbul, TR

Abstract In this study, application of the steam injection method (SIM), Miller cycle (MC) and turbo charging (TC) techniques into a four stroke, direct-injection diesel engine has been numerically and empirically conducted. NOx emissions have detrimental influences on the environment and living beings. They are formed at the high temperatures, thus the Diesel engines are serious NOx generation sources since they have higher compression ratios and higher combustion temperatures. The international regulations have decreased the emission limits due to environmental reasons. The Miller cycle (MC) application and steam injection method (SIM) have been popular to abate NOx produced from the internal combustion engines (ICEs), in the recent years. However, the MC application can cause a reduction in power output. The most known technique which maximizes the engine power and abates exhaust emissions is TC. Therefore, if these three techniques are combined, the power loss can be tolerated and pollutant emissions can be minimized. While the application of the MC and SIM causes to diminish in the brake power and brake thermal efficiency of the engine up to 6.5% and 10%, the TC increases the brake power and brake thermal efficiency of the engine up to 18% and 12%. The experimental and theoretical results have been compared in terms of the torque, the specific fuel consumption (SFC), the brake power and the brake thermal efficiency. The results acquired from theoretical modeling have been validated with empirical data with less than 7% maximum error. The results showed that developed

1

*Corresponding Author. Tel: 90 212 383 2950 Fax: +90 212 383 2941, e-mail: [email protected]

1

combination can increase the engine performance and the method can be easily applied to the Diesel engines. Keywords: Finite-time thermodynamics; Miller cycle, Steam injection method; Turbo charging. 1. Introduction The environmental regulations and restrictions force to decrease emissions released from ICEs. One of the most detrimental emissions is NOx and it is formed at high combustion temperatures. The prior NOx reduction methods can decrease the engine performance. The problem is to reduce NOx emissions without any performance loss. So many studies have been carried out and new methods have been developed to decrease NOx emissions. One of these techniques is the SIM which can provide high decrease rates in the NOx formation [112]. Another NOx control technique, which has been widespread in the recent years, is the MC application. On the other hand, this method can lead to minimization in the engine power [13-23]. TC method is commonly preferred to increase specific power of the ICEs [40-51]. These three methods could be combined so as to provide the highest NOx abatement rate with higher power output for the diesel engines. The dual and triple combinations are presented as MC-TC, MC-SIM and MC-SIM-TC, respectively, in this study. The SIM was firstly proposed by Parlak et al. [1]. The proposed technique lowered NOx formation up to 33% in comparison with the standard (STD) condition of a diesel engine. Also, the performance characteristics improved and the brake thermal efficiency enhanced by 3%. Kokkulunk et al. [2-3] combined the SIM and exhaust gas recirculation (EGR) to decrease NOx formation and to improve brake efficiency and power output characteristics of a diesel engine. Gonca et al. [4-5] performed a study to compare a diesel engine with SIM and the MCDE. They reported that the maximum reduction in NO formation was observed at the combination of the SIM and MC applications. Cesur et al. [6] and Kokkulunk et al. [7] 2

applied the SIM into spark ignition and compressed ignition engines, a diminishment in NOx formation and performance improvement was observed. Gonca et al. [8] performed an investigation to define the optimal mass ratios and temperatures of vaporized water for ICEs with SIM and TC applications. Gonca [9-10] carried out the SIM application for a diesel engine operating with the ethanol-diesel mixtures [9] and operating with hydrogen enriched air [10]. Parlak et al. [11] empirically applied the SIM into the intake manifold of a diesel engine run on tobacco seed oil-based biodiesel. Gonca [12] performed an investigation and comparison study on the influences of the SIM on thermodynamic characteristics of the biofuels. The used fuels were classified as biodiesels and alcohols in terms of the effect of the SIM. It was reported that higher NO reduction was seen in alcohols compared to biodiesels. The application of the MC into the ICEs minimizes the NOx at very high ratios. Hence, so many studies have been done recently. Wang et al. [18] showed a considerable diminishment in NOx formation of a diesel engine with the experimental application of the MC. Wang et al. [19-20] performed a numerical study [19] and experiments [20] for the MC SI Engine so as to abate NOx formation. Mikalsen et al. [21] applied the MC into a gasoline engine operating with natural gas. Gonca et al. [22,23] proved that the MCDE has less NO formation and higher brake thermal efficiency compared to the standard diesel engine. Al-Sarkhi et al. [2426] examined the impacts of the temperature-dependent specific heats on the performance characteristics for the reversible MC [24] and irreversible MC [25]. Also, Al-Sarkhi et al. [26] investigated the variation of the power density depending on engine design parameters for the MC. Zhao and Chen [27] analyzed the cycle performance characteristics of the irreversible MC depending on pressure ratio change. Ebrahimi [28-29] used the finite-time thermodynamics modeling (FTTM) to analyze the performance variation of the reversible and irreversible MC engines considering engine speed variation and temperature dependent specific heats [28] and considering the variation of stroke length and air/fuel ratio [29]. 3

Rinaldini et al. [30] abated NOx and soot emission produced from a high speed diesel engine with the MC application. Li et al. [31] conducted an empirical examination to show the influences of the MC on the SFC of a spark ignition engine. Wu et al. [32] applied the MC into a supercharged SI engine.

4

Application of the TC into the ICEs maximizes the performance. Therefore, so many applications have been performed related to turbo charging systems. Can et al. [52] investigated the influences of ethanol-diesel mixtures and fuel injection pressures on the emission formation and the performance specifications of a TC diesel engine. Rakopoulos and Giakoumis [53] performed a study to analysis the exergy characteristics of a TC diesel engine. 5

Cinar et al. [54] performed an investigation to understand the impact of using CO 2 addition into the intake manifold and pressure of the injection on the performance specifications and emission formations of a indirect injection diesel engine with TC application. Tauzia et al. [55] presented a new model to predict heat release rate curve for a modern TC diesel engine. Giakoumis [56] examined the influences of cylinder wall insulation on the performance characteristics of a TC diesel engine in terms of the first law and second law of the thermodynamics. Giakoumis [57] carried out a study to examine the influences of the lubricating oil properties on the transient response of a TDE. Rakopoulos et al. [58] conducted an empirical work so as to demonstrate the mechanisms of smoke and NO formations for TC diesel engine running on butanol-diesel and biodiesel-diesel mixtures under various transient conditions. Rakopoulos et al. [59] carried out an empirical work to determine the mechanisms of smoke, NO and combustion noise formations during hot starting for biodiesel-diesel and butanol-diesel mixtures. Giakoumis et al. [60] performed various experiments to understand the characteristics of combustion noise emission of a TC diesel engine. Chiong et al. [61] studied on the performance specifications of an engine which has turbocharger. Liang et al. [62] analyzed the impact of oxygen enrichment of the intake air and water/diesel emulsion on the combustion characteristics of a TC diesel engine. Cornolti et al. [63] performed flow analysis in the intake and exhaust manifolds of a TC diesel engine with different EGR application methods. Gonca et al. [64] applied the TC method into a MCDE using a screw type compressor to increase engine performance and decrease NO formation simultaneously. Gonca and Sahin [65] showed that engine design and operating parameters could be optimized to increase the performance of a TC diesel engine with the MC and SIM applications. They used a realistic FTTM which is comparable with the empirical data. This study presents the influences of the combination of the SIM, MC and TC methods (SIMMC-TC) on the torque, the brake power, the brake thermal efficiency and SFC of a direct 6

injection diesel engine. The torque, brake power, brake efficiency and SFC have been empirically acquired. The empirical data have been compared with the results of a FTTM and maximum error obtained is less than 7%. Apart from prior studies, this study also presents a comprehensive comparison of the results of experiments and the FTMM. Also, the results of this study showed that the performance can increase with the developed combination. 2. Materials and methods 2.1. Experimental set-up The experimental studies were carried out with a single-cylinder diesel engine. Table 1. demonstrates the engine specifications. The MC is obtained by 10 CA retardation of intake valve (IV) closing. The original cam-shaft is cam-shaft 52 (STD) which means the intake valve is closed 52 CA after the bottom dead center during the compression process. 10 CA retardation is carried out with cams-haft 62 (C62). In order to implement the MC into the engine, the modified camshaft was input into the engine used. The pictures and technical drawings of the cam shafts are presented in the Fig. 1. The electronically controlled SIM was applied using a boiler with electric resistance. The pressure and temperature conditions of the steam are 3 bar and 133.5 oC in the boiler The optimum steam ratio has been found as 20% steam ratio of fuel mass in terms of power output, fuel economy and NOx reduction in the previous studies [1,4,6,7,11,13,70], thus, this ratio is used in this study for the steam injected conditions. The turbo charging has been applied into the engine using a compressor removed from E 211 brand Mercedes-Benz truck. The compressor has been input to intake line of the engine and 1.1 bar (T1.1) turbo charging pressure has been provided by an electronic control unit. The empirical set-up is shown in the Fig. 2. A dynamometer which has 20 kW absorbing capacity and “S” type load cell with the precision of 0.01 kg were used so as to measure engine torque. The experiments were carried 7

out between 1500 and 3000 rpm engine speeds and at full load conditions. Firstly, standard cam-shaft was operated at the naturally aspirated condition and then the other cam-shaft which provides the MC was input into the engine at 20% steam injection ratio and 1.1 bar TC pressure conditions. The empirical works were carried out three times for each condition and then performance characteristics were compared with those of STD condition. The uncertainties for the measurements are given in the Table 2. 2.2. Theoretical model A new FTTM is used to obtain theoretical results [65]. The model takes friction, variable specific heats, heat transfer and combustion effects into account. In the model used, the specific heats depending on temperature variation for constant pressure and constant volume are expressed as below [66]: CP  2.506 1011T 2  1.454 107 T 1.5  4.246 107 T  3.162 105 T 0.5 1.3301  1.512 104 T 1.5  3.063 105 T 2  2.212 107 T 3

CV  CP  R

(1)

(2)

Where R is the gas constant. In order to attain reasonable results, Dual-Miller cycle (DMC) is used in the simulation model. The cycle is demonstrated in the Fig.3. The total heat input Qin at the process (2-3) and at the process (3-4) could be written as below:

8

T3

T4

T2

T3

Qin  mt  CV ,mix dT  mt  CP ,mix dT 

mt

3     T 2.5 T2 T 1.5 11 T  1.454 107  4.246 107  3.162 10 5      2.506 10 3 2.5 2 1.5 m    0.5  a    T T 2    4 5 1 7  1.0433T  1.512 10     3.063 10  T   2.212 10      0.5   2      3   T 2.5 T2 T 1.5 11 T   1.454 107  4.246 107  3.162 105    2.506 10 3 2.5 2 1.5     2ms  0.5 2    4 T 5 1 7 T  1.0433T  1.512 10   0.5   3.063 10  T   2.212 10   2           

+ mt

T3

T2

ma  ms 3     T 2.5 T2 T 1.5 11 T  1.454 107  4.246 107  3.162 105      2.506 10 3 2.5 2 1.5 m    0.5  a    T T 2    4 5 1 7   1.3301T  1.512 10     3.063 10  T   2.212 10      0.5   2      3   T 2.5 T2 T 1.5 11 T   1.454 107  4.246 107  3.162 105    2.506 10 3 2.5 2 1.5     2ms  0.5 2      T T 4 5 1 7  1.3301T  1.512 10   0.5   3.063 10  T   2.212 10   2           

T4

(3)

T3

ma  ms

Where CV ,mix , CP ,mix and Rmix are the constant volume specific heat, constant pressure specific heat and gas constant of the steam-air mixture. They are written as follows: CV ,mix 

CV ,a ma  CV , s ms ma  ms

(4)

CP,mix  CV ,mix  Rmix

(5)

Ra ma  Rs ms ma  ms

(6)

Rmix 

Where CV , s and CV ,a are the constant volume specific heats of the steam and air. The relation between them can be given as below:

CV , s  2CV ,a

(7)

Rs and Ra are the gas constant of the steam and air. Their values are given as 0.4615 kJ/kg.K and 0.287kJ/kg.K, respectively. ma , mt and ms are masses of the net intake air, total charge and steam per second, respectively. They can written as below:

9

ma N m f Fst ,  120 

(8)

 F  mt  m f 1  st   ms ,   

(9)

ma 

ms 

ms N Vste s N   xm f 120 120

(10)

Where N (rpm) is engine speed. It is used for conversion of mass rate per second from mass per cycle.  , x , Fst ,  s and Vste are equivalence ratio, steam percentage compared to fuel mass, stoichiometric fuel-air ratio, steam volume, steam density which are given as follows: m  Fst   f   ma  st

mf 

(11)

mf N

(12)

120

msuc  VT a , ma  a (

(13)

VT  Vste )+ s Vste , rM

ms  xm f , Vste 

(15)

Rs x aVT  , Fs Ra a  x a Rs

 s  f (Tmix , P1 ) 

(14)

Ra a Rs

(16)

(17)

mf



ma Fst

(18)

10

f (...) stand for functional expressions. The functional values are attained from EES software [67]. Where  a and Tmix are average air density and temperature of air-steam mixture which are given as follows: Tmix 

ma ,iTa ,i Ra  msTs Rs

(19)

ma ,i Ra  ms Rs

a  f (Tmix , P1 )

(20)

Where subscript "i" means the state before the starting of vapor injection and it can be called as the initial condition. Ts and Ta ,i are the steam temperature and dry air temperature. Ts is 133 o

C. ma ,i is the mass rate of dry air at the initial condition and it can be written as:

ma ,i  a ,i (Vs +Vc )=a ,i VT  a ,i

(Vs r ) r 1

(21)

Where ,  a ,i , Vc , Vs and VT are dry air density, clearance volume, stroke volume and total volume which are expressed as:

Vs   r  1

VT r

Vc  VT  Vs 

(22) VT  d 2 L 1  r 4 r 1

a,i  f (Ti , P1 )

(23) (24)

In the open literature, the compression ratio ( r ) is known as:

r  V1 / V2

(25)

Isentropic efficiencies for compression process and expansion process are determined as below [68]:

C 

T2S  T1 T2  T1

(26)

and 11

E 

T4  T5 T4  T5S

(27)

Several dimensionless engine design parameters such as cycle temperature ratio (  ),cycle pressure ratio (  ),cut-off ratio ( ), pressure ratio (  ), miller cycle ratio ( rM ),expansion ratio (  ) and stroke ratio (  ) are determined as below:



 r k 1  1  Tmax T4 r       m   1  Tmin T1 r   C  

(28)

  Pmax / Pmin  P3 / P1  P4 / P1

(29)

  v4 / v3  T4 / T3

(30)

  P3 / P2  T3 / T2 .

(31)

rM  V6 / V1  T6 / T1

(32)



rM r





 . 

(33)

The total heat rejection Qout at constant volume (5-6) and pressure (6-1) may be given as below:

12

 m  msuc  T5 Qout   t  T6 CV ,mix dT   msuc  ma  CP ,T6 T6  2  

 mt  msuc  2 

3     T 2.5 T2 T 1.5 11 T  1.454 107  4.246 107  3.162 105      2.506 10 3 2.5 2 1.5 m    0.5  a    T T 2    4 5 1 7  1.0433T  1.512 10     3.063 10  T   2.212 10      0.5   2      3   T 2.5 T2 T 1.5 11 T   1.454 107  4.246 107  3.162 105    2.506 10 3 2.5 2 1.5     2ms  0.5 2      T T 4 5 1 7  1.0433T  1.512 10   0.5   3.063 10  T   2.212 10   2              ma  ms 

T5

(34) T6

  2.506 1011T6 2  1.454 107 T61.5  4.246 107 T6  3.162 105 T6 0.5     ma     4 1.5 5 2 7 3   1.3301  1.512 10 T6  3.063 10 T6  2.212 10 T6     11 2 7 1.5 7 5 0.5  2.506 10 T6  1.454 10 T6  4.246 10 T6  3.162 10 T6        2ms 1.3301  1.512 104 T 1.5  3.063 105 T 2  2.212 107 T 3 6 6 6    +  msuc  ma  T6 ma  ms

Where msuc suction air mass rate per second. It can written as below: msuc 

msuc N , 120

(35)

Where msuc suction air mass per cycle. The process (1-2s) and process (4-5s) can be stated as [68]:

CV1 ,mix  ln

T2 s T   Rmix ln r , CV ,mix  ln 5 s  Rmix ln T1 T4 r  rM 2

(36)

Where CV1 ,mix 

CV1 ma  CV1 , s ms ma  ms

, CV2 ,mix 

CV2 ma  CV2 , s ms ma  ms

CV1 , s  2CV1 , CV2 ,s  2CV2

CV1  2.506 1011T2 s12  1.454 107 T2 s11.5  4.246 107 T2 s1  3.162 105 T2 s10.5  1.0433  1.512 104 T2 s11.5  3.063 105 T2 s12  2.212 107 T2 s13

CV2  2.506 1011T5 s 4 2  1.454 107 T5 s 41.5  4.246 107 T5 s 4  3.162 105 T5 s 4 0.5  1.0433  1.512 104 T5 s 4 1.5  3.063 105 T5 s 4 2  2.212 107 T5 s 4 3

(37)

(38)

(39)

(40)

13

T2 s1 

T2 s  T1 T T , T5 s 4  5 s 4 T T ln 5 s ln 2 s T4 T1

(41)

The brake thermal efficiency and brake power are expressed as:

br 

Pbr , Pbr  Qin  Qout  Pl Qf

(42)

Where Pl is loss power depending on friction which is:

Pl   SP 2

(43)

where  is friction coefficient, S P is average velocity of the piston which is:

Sp 

L N 30

(44)

Where L is stroke height. Q f is the total energy content of the fuel which is: Q f  m f Hu

(45)

Where H u is lower heat value of the injected fuel. The total heat input in the cycle: Qin  Q f ,c  Qht

Where Q f ,c is heat released by combustion; Qht is the loss energy by heat transfer they can be stated as: Q f ,c  c m f H u

Qht  htr Acyl (Tme  TW )=htr Acyl (

(46) T2  T4  TW ) 2

(47)

Where c is combustion efficiency which was:

c  0,99743  4,783  105 N  0,0768868x  0,0014086Pa  0,0004245rM  0,31948

(48)

14

Where Pa , rM , htr are intake pressure, the intake valve retarding angle, heat transfer coefficient. Acyl is the area in which heat transfer carried out which can be written as: r d2 Acyl   dL  r 1 2

(49)

Where Tme , TW and d are average combustion temperature, cylinder wall temperature and cylinder bore. htr is calculated using Hohenberg correlation [69]: htr  130VT 0.06 P10.8Tmix 0.4 (Sp +1.4)0.8

(50)

The simulation results and experimental data have been compared and an acceptable approximation was attained.

3. Results and discussion

Fig. 4-5 show the torque and brake power depending on the engine speed for different engine modes. The engine speeds change from 1500 rpm to 3000 rpm. The torque and brake power are between 25-45 Nm and 5-10 kW, respectively. It is obvious that the torque abates and brake power enhances with rising engine speed. While lower in-cylinder pressures lead to lower engine torque, higher engine speeds provides higher engine power. Therefore, as the engine torque decreases, the engine power raises. The results showed that low engine speeds modes should be selected for the higher torque requirement, high engine speeds should be selected for the higher power requirement and medium engine speeds should be selected for the lower SFC requirement and higher fuel economy. Equivalence ratio affects the general performance characteristics of the engine since maximum combustion temperatures are observed about 1 of the equivalence ratio. The performance characteristics are deteriorated at the higher and lower equivalence ratio. The 15

volumetric efficiency affects the intake air mass and total charge in the cylinder. Therefore, the performance characteristics improve as

volumetric efficiency increases.

The

implementation of the SIM and the MC into the used engine abates the torque and brake power due to higher equivalence ratio and lower volumetric efficiency. However, the application of the TC increases the torque and brake power owing to lower equivalence ratio and higher oxygen concentration depending on higher volumetric efficiency. Minimized equivalence ratio and maximized oxygen concentration provides higher brake thermal efficiency. Also, combustion efficiency enhances with increasing oxygen concentration. Therefore, the engine torque and power increase. If the results are investigated, the maximum and minimum decrease rates in the torque are seen at C62-S20 condition as 6.5% (at 3000 rpm) and 3% (at 1500 rpm). The maximum and minimum increase rates in the torque are 18% with 1500 rpm at C62-T1.1 and 10.4% with 2100 rpm at C62-S20-T1.1. The lowest torque and brake power are seen with C62-S20 at 3000 rpm and 1500 rpm, as 25.4 Nm and 5.4 kW. The maximum torque and EP are seen with C62-T1.1 at 1500 rpm and 3000 rpm, as 41.2 Nm and 9.7 kW. The evaluation of the brake thermal efficiency and SFC depending on the engine speed for different engine modes is demonstrated in the Fig. 6-7. The maximum brake thermal efficiency and the minimum SFC are obtained at the medium engine speeds. Because, at the higher and lower engine speeds, the losses depending on friction, heat transfer and incomplete combustion is higher. Therefore, the brake thermal efficiency decreases and SFC increases at the lower and higher engine speeds. The application of the MC and SIM abates the brake thermal and SFC due to lower volumetric efficiency. On the other hand, TC increases the brake thermal efficiency and decreases the SFC. When the obtained results are examined, it is observed that the maximum decrease rate in the brake thermal efficiency is 10% and it is seen at C62-S20 and 2400 rpm. The lowest decrease rate is seen at 1800 rpm, as 6.5%. The highest 16

increment rate is acquired with C62-T1.1, as 12% at 1500 rpm, the lowest increment rate is seen with C62-S20-T1.1, as 4% at 2400 rpm. The lowest brake thermal efficiency and highest SFC are reached with C62-S20, as 24.3% and 353 g/kWh at 3000 rpm; the maximum brake thermal efficiency and minimum SFC are reached with C62-T1.1, as 33.3% and 257 g/kWh at 2100 rpm. It is obvious from the figures that the results obtained with the mathematical model are close those obtained with experimental study. The maximum error is less than 7% and it is seen with C62 condition at 2100 rpm. The obtained results were supported by the previous studies [1-4, 6,7,11] in terms of the effects of the SIM on the performance of the ICEs. The MC application was also used in the previous works [18-22, 30,31,34, 64]. They reported the similar results to the acquired results presented in this study. The TC application is known as power increment method in the literature [55, 56, 64, 65]. As expected, the brake power of the test engine decreased by the application of TC up to 18%.

4. Conclusion The presented work reported the influences of the MC, the SIM and TC applications on the performance characteristics of a diesel engine. The torque, brake power output and brake thermal efficiency remarkably decreased by applying the MC and SIM, however they are improved by TC. The maximum performance is obtained at the C62-T1.1 condition. The highest increase rates in the brake power and brake thermal efficiency are 18% and 12%. The minimum performance is seen at the C62-S20 condition. The highest reduction rates in the brake power and brake thermal efficiency are 6.5% and 10%. It should be pointed out that the modes with turbo charging method should be chosen to increase higher engine performance. In this study, a theoretical study has been carried out based on the FTTM. The results obtained have been validated with non-noticeable difference in terms of torque, brake power, brake thermal efficiency and SFC. The combination of MC, SIM and TC is novel and unique 17

and it can be applied to the Diesel engines to increase engine performance at the lower emission formations.

18

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Figure Captions Fig. 1. The drawings and picture of the original and modified cam-shaft. Fig. 2. Experimental set-up. Fig. 3. P-v and T-s schematic diagrams for DMC [23]. Fig. 4. Comparison of empirical data and theoretical results for torque Fig. 5. Comparison of empirical data and theoretical results for brake power. Fig. 6. Comparison of empirical data and theoretical results for brake thermal efficiency. Fig. 7. Comparison of empirical data and theoretical results for SFC.

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TABLES Table1. Engine properties Engine type

Antor

Bore [mm] Stroke [mm]

85 90

Cylinder Number

1 3

Stroke Volume [dm ]

0.51

Power, 2700 rpm [kW]

9

Injection pressure [bar]

175

Injecting timing [crank angle]

28

Compression ratio

17.5

Maximum speed [rpm]

3000

Cooling

Air

Injection

Direct injection

Table2. The errors in parameters and total uncertainties Parameters

Systematic errors 

Load [N] Speed [rpm]

0.1 1.0

Time [s]

0.1

Temperature [oC]

1.0

Fuel consumption [g]

0.01

SFC [g/kWh]

1.5

Torque [Nm]

1.1

28

Figures

a) Original camshaft [34]

b) Camshaft C62 (10 CA retarding) [34]

Fig. 1.

29

Fig. 2.

30

P

T 3

4

 Q 34

2 2s

 Q 23

4

3

5 5s

2 5 5s

2s 1

1

6 V

6

Q56

Q61 S

Fig. 3.

31

Exp-STD Exp-C62 Exp-C62-T1.1 Exp-C62-S20 Exp-C62-S20-T1.1

Torque, Nm

40

Mod-STD Mod-C62 Mod-C62-T1.1 Mod-C62-S20 Mod-C62-S20-T1.1

35

30

25 1500

1800

2100 2400 Engine Speed, rpm

2700

3000

Fig. 4.

32

10

Brake Power, kW

9

8

7

Exp-STD Exp-C62 Exp-C62-T1.1 Exp-C62-S20 Exp-C62-S20-T1.1

6

5 1500

1800

2100 2400 Motor Devri, d/d

Mod-STD Mod-C62 Mod-C62-T1.1 Mod-C62-S20 Mod-C62-S20-T1.1

2700

3000

Fig. 5.

33

34

Brake Thermal Efficiency, %

32

30

28

Exp-STD Exp-C62 Exp-C62-T1.1 Exp-C62-S20 Exp-C62-S20-T1.1

26

24 1500

1800

Mod-STD Mod-C62 Mod-C62-T1.1 Mod-C62-S20 Mod-C62-S20-T1.1

2100 2400 Motor Devri, d/d

2700

3000

Fig. 6.

34

Exp-STD Exp-C62 Exp-C62-T1.1 Exp-C62-S20 Exp-C62-S20-T1.1

350

SFC, g/kWh

330

Mod-STD Mod-C62 Mod-C62-T1.1 Mod-C62-S20 Mod-C62-S20-T1.1

310

290

270

250 1500

1800

2100 2400 Engine Speed, rpm

2700

3000

Fig. 7.

35

STEAM INJECTION

STANDARD CAMSHAFT

MODIFIED CAMSHAFT

HIGHER NOX

LOWER NOX

36

Highlights  Performance of a diesel engine is simulated by finite time thermodynamics.  Effect of steam injection on performance of a Miller cycle engine is examined.  Model results are verified with the experimental data with less than 7% error.

37