Cycle Simulation

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performance of a diesel engine fuelled by diesel and also the

different blends of diesel and biodiesel. The model predicts the

performance of a CI engine in terms of brake power and brake

thermal efficiency for all the fuels considered for the present study.

Fuel properties [11] and the engine design and operating parame-

ters are specified as inputs to the model.

2. General description of the model

It is assumed that there is spatial uniformity of pressure, temper-

ature andcomposition of thecylindercontentat eachcrank angle.Thefuels considered are diesel and 20% (B20), 40% (B40), 60% (B60) (by

volume) blending of biodieselfrom Karanjaoil [11,14] with diesel. The

molecularformula of diesel is approximated as C12H23 [15]. Karanjaoil

is mostly oleic and linoleic. The chemical formula of pure biodiesel is

considered as C17.75H33.43O1.98 and this has been derived from the

fatty acid composition of karanja oil as given in Table 1. It is also

assumed that the air fuel mixture is lean and this leads to tempera-

tures at which dissociation of products does not have much effect on

engine performance [16]. Therefore, dissociation of products of

combustion is neglected in order to keep the analysis simple [17,18].

However, combustion products should be defined by considering

dissociation in order to represent a more realistic cycle [17,19]. Poly-

nomial expressions are used for each species (O2, N2, CO2, H2O)

considered in the calculation of specific heats, internal energy andenthalpy as a function of temperature [13,19–21]. The compression

phase begins at the point of closing of the inlet valve (IVC) and

continues up-to crank angle at which combustion begins. The period

from the endof combustion tothe exhaust valve opening (EVO) isthe

expansion phase. The compression and the expansion phases are

considered as polytropic.

2.1. Energy conservation

The first law of thermodynamics applied for the closed cycle

period (from IVC to EVO) can be written as

dQ ndq

¼ dU

dq þ dW

dq (1)

Where dQ ndq

is the net heat release rate and is the difference between

dQ c

dq and dQ h

dq

dQ c

dq ¼ Heat release rate due to combustion of fuel

dQ hdq

¼ Rate of heat transfer from in cylinder gases to the wall

dU

dq ¼ Rate of change of internal energy

dW

dq ¼ Rate of work transfer

Equation (1) can be rearranged as

dQ c

dq dQ h

dq ¼ MC v

dT

dqþ p

dV

dq (2)

M , C v, p, T and V are the mass, specific heat at constant volume,

instantaneous pressure, instantaneous temperature, instantaneous

volume of the cylinder content respectively.

Instantaneous cylinder volume V is given by,

A ¼ 2:94Dv

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiL2v LvDv

8 þ D2

v

128

!v uut (3)

The terms used in equation (3) are as given below.

Table 1

Fatty acid composition of karanja oil [14].

Sl. No Fat ty acid Mo lecular formula Co mpo sition(wt.%)

1 Palmitic C16H32O2 11.30

2 Stearic C18H36O2 9.80

3 Oleic C18H34O2 45.25

4 linoleic C18H32O2 24.75

5 Srachidic C20H40O2 1.75

6 linolenic C18H30O2 2.90

7 behenic C22H44O2 3.20

8 Unidentified 1.05

Nomenclature

A Area, m2

C v Specific heat at constant volume, kJ/kmol

D Cylinder bore, mD p Valve port diameter, mm

Ds Valve stem diameter, mm

Dv valve diameter, mm

Dvi Inlet valve diameter, mmk thermal conductivity, W/mK

L Length of connecting rod, mm

Lv Instantaneous valve lift, mm

M Mass, kgN Engine RPM

n no. of intake valve/cylinder

n pr no. of piston ring

p Pressure, barP sl Piston skirt length, mm

pe Exhaust back pressure, bar

pimf Inlet manifold pressure, bar

pa atmospheric pressure, bar

pimep indicated mean effective pressure, bar

dQ c

dq Heat release rate due to combustion of fuel, kJ/degree

crank angledQ hdq

Rate of heat transfer from in cylinder gases to the wall,

kJ/degree crank angledQ ndq

Net heat release rate, kJ/degree crank angleR Universal gas constant, kJ/molK

Re Reynolds’s Number

S stroke length, mm

T temperature, Kt time, secdU dq

Rate of change of internal energy, kJ/degree crank

angle

V Volume, m3

V s Stroke volume, m3

V p Mean piston speed, m/sdW dq

Rate of Work transfer, kJ/degree crank angle

Greek Letters

r density, kg/m3

m dynamic viscosity, kg/m sec

g Ratio of specific heat, dimensionless

q Crank angle, degreef fuel air equivalence ratio, dimensionless

T.K. Gogoi, D.C. Baruah / Energy 35 (2010) 1317–13231318



Rc is the compression ratio, L is the length of connecting rod, S is

the stroke length,

V s is the stroke volume of the cylinder and q is the crank angle.

2.2. Ignition delay

The ignition delay is the time between start of fuel injection and

start of combustion. Many expressions for ignition delay are found

in literature as a function of mixture pressure, temperature, fuel

cetane number and fuel air equivalence ratio. The following

empirical correlation is used for calculation of ignition delay

[16,21,22].

sid ¼ 2:64

p0:8f0:2exp

16550 20CN

RT

(4)

Where R is universal gas constant, CN is fuel cetane numberand f is

fuel air equivalence ratio. However, the constants and exponents in

the above correlation are to be better calibrated against experi-

mental results.

2.3. Combustion

In the present work, the combustion process is modeled

using a single zone approach, which is based on a uniformly

distributed heat releasing phenomenon. Spray combustion is not

considered in detail. Detailed phenomenological combustion

modeling is undoubtedly useful, particularly in the design of

combustion chamber; however, single zone heat release

combustion models are appropriate for use in total diesel system

simulation where the combustion process details are not the

primary focus [13].

The Weibe function is used for calculation of heat release rate

due to combustion of fuel.

x ¼ 1 exp" a

q q0Dq

m

þ1#

(5)

Where x is the mass fraction burned, q0 is the start of combustion

and Dq is the combustion duration. a is an adjustable parameter

that characterizes the completeness of combustion. The parameter

m represents the rate of combustion. Differentiating x with respect

to qq0

Dq and multiplying withQ av

Dq, the heat release rate is calculated

as,

dQ c

dq ¼ aðm þ 1Þ

Q avDq

q q0

Dq

m

exp

" a

q q0

Dq

mþ1#

(6)

Where Q av is the heat released per cycle. The value of m for all the

fuel is taken as 2.0 and the value of a for diesel, B20, B40 and B60

were taken as 5.0 [13], 5.508, 6.008 and 6.508 respectively.

2.4. Heat transfer

The heat transfer between the trapped gas and the surrounding

wall is calculated by using Anand’s equation.

dQ h=dq

A ¼ a

k

DRebðT w T Þ þ c

T 4w T 4

(7)

Where A is the heat transfer area and T w is the temperature of the

cylinder wall. Re ¼ rV pDm is the Reynolds number with D the

cylinder bore and V p the mean piston speed. r, m and k are the gas

density, dynamic viscosity and thermal conductivity respectively.

The value of a varies with speed and engine design. With normal

combustion, 0:35 a 0:8 with b ¼ 0.7 [13]; c ¼ 0 for the

compression period and otherwise c ¼ 3.3 108 W/m2K4 is the

most usual value [5].

2.5. Gas exchange process

During intake and exhaust processes, the cylinder pressure is

not only a function of in-cylinder processes but also of mass flowrate through the valves. Applying energy equation to these

processes and treating the gas as ideal [17,20,21,23], one obtains,

For the exhaust process,

dp

dt ¼ g p

1

M

dM

dt 1

V

dV

dt

exhaust

(8)

For the intake process,

dp

dt ¼ g

RT

V

dM

dt p

V

dV

dt

intake

(9)

Both these equations require mass flow rate dM dt

, which can be

determined from the equation given below [21].

dM

dt ¼ Ap0

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2g

RT ðg 1Þ

p

p0

gg1

p

p0

gg1

1

s (10)

Where A is the instantaneous valve flow area and it depends upon

the valve lift and the geometric features of the valve head, seat and

stem. Depending upon the instantaneous valve lift, three different

cases may arise [24],

CaseI :0 < Lv

Dv

< 0:125: The minimum flow area corresponds to

the slant surface of a frustum of a right circular cone and perpen-

dicular to the valve seat.

A ¼ 2:22Lv

7

8Dv þ Lv

2

(11)

Where Lv and Dv are the instantaneous valve lift and valve headdiameter respectively.

CaseII : 0:125 < Lv

Dv

< 0:2735: The minimum flow area is still the

slant surface but no longer perpendicular to the valve seat.

V ðqÞ ¼ V s

24 Rc

Rc 1 1 cosq

2 þ L

S 1

2

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi"2

L

S

2

sin2q

#v uut35 (12)

CaseIII : Lv

Dv

> 0:2735: When valve lift becomes large, then the

flow area is port flow area minus the sectional area of the valve

stem.

A ¼ p

4

"

Dv

2 2

Dv

4 2#

(13)

It is assumed that the port diameterD p ¼ Dv

2 and the valve stem

diameterDs ¼ Dv

4 , with a valve seat angle of 45.

2.6. Net work

Net work done in a complete cycle is given by

W net ¼ #

p þ D p

2

DV (14)

Where D p is the change in pressure inside the cylinder due to

piston motion, combustion, heat transfer and flow into and out of

the cylinder. The following equation gives the pressure drop due to

heat transfer [21].

T.K. Gogoi, D.C. Baruah / Energy 35 (2010) 1317–1323 1319



D p

p ¼

dQ hdt

MC vT Dt (15)

2.7. Frictional power

Frictional losses affect the maximum brake torque and the

minimum brake specific fuel consumption directly and are often

a criterion of good engine design. These losses not only reduce the

power but also influence the size of the coolant systems [16]. The

mean effective losses of power due to friction in different moving

parts are calculated by using the following empirical relations

[21].

(i) Mean effective pressure(FMEP)lost due to friction in the piston

and piston rings

FMEP 1 ¼ 12:85P sl

DL 100 V p

1000 (16)

Where P sl is the piston skirt length (mm)

D is the cylinder bore (m), L is the stroke length (m), V p is mean

piston speed (m/s)

(i) MEP lost in bearing friction

FMEP 2 ¼ 0:0564D

L N

1000 (17)

N is the engine RPM.

(iii) MEP lost in friction in the valve gear

FMEP 3 ¼ 0:226

30 4N

1000

nD1:75

vi

D2L (18)

Where n is the number of intake valve/cylinder, Dvi is the inlet valve

diameter.

(iv) MEP lost in overcoming inlet and throttling losses

FMEP 4 ¼ pe

2:75þ pimf (19)

Where pe is the exhaust gas back pressure (bar) and pimf is the inlet

manifold pressure (bar).

(v) MEP lost in pumping

FMEP 5 ¼ 0:0275

N

1000

1:5

(20)

(vi) MEP lost in friction due to gas pressure behind rings

FMEP 6 ¼ 0:42

pa pimf

L

D2

0:0888Rc þ 0:182R1:330:394V p=100c

10 ð21Þ

Where pa is the atmospheric pressure (bar).

(vii) MEP lost in friction due to wall tension in rings

FMEP 7 ¼ 10 0:377Ln pr

D2 (22)

Where n pr is the number of piston rings.

(viii) Blow by losses

FMEP 8 ¼ ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi pa pimf

p "

0:121R0:4c ð0:0345 þ 0:001055Rc Þ

N

100

1:185#

(23)

(ix) MEP lost in overcoming combustion chamber and wall

pumping losses

FMEP 9 ¼ ffiffiffiffiffiffiffiffiffiffiffiffiffi

pimep

11:45

r 0:0915

N

1000

1:7

(24)

Where pimep is the indicated mean effective pressure.

Indicated power can be calculated from the net work done

during the cycle and hence the indicated mean effective pressure.

The brake mean effective pressure is the difference between the

indicated mean effective pressure and the frictional mean effective

pressure. Thus, it calculates brake power from known values of

brake mean effective pressure, stroke volume and speed.

3. Results and discussion

The developed model was used to investigate the effect of

variation of speed and compression ratio (CR) on engine perfor-

mance (Brake Power and brake thermal efficiency) fuelled by diesel

and blends of diesel and biodiesel. Table 2 shows the engine

configuration and the engine operating conditions. The properties

of the fuels, viz. diesel and the various blendsof diesel and biodiesel

(B20, B40 and B60) are shown in Table 3. The speed was varied from

800 to 1800 rpm and the CR from 12 to 18.5.

3.1. Effect of speed on brake power

Figs. 1–3 show the variations of brake power with speed for CR 14, 15.5 and 17.5. The results show that bake power increases with

the increase in speed for all the fuel, the peak power occurs at

a particular speed and further increase in speed results in decrease

of brake power. Peak power at a particular speed is a characteristic

of diesel engine. At speeds above the one at which peak power

occurs, the frictional losses increase very rapidly and hence the

brake power decreases. From the figures, increased brake power is

evident in case of the blends B40 and B60. The brake power output

in case of the blend B20 is slightly lower than diesel at all speeds

and CR. Rehman et al. [12] also on a variable speed TD43F engine,

obtained higher performance with karanja methyl ester and its

blends B40 and B60 and lower brake power in case of B20.

According to their report, the increased power in case of the blends

Table 2

Engine design and operating parameters.

Parameters Value

Bore, mm 87.5

Stroke, mm 110

Connecting rod length, mm 230

Inlet valve diameter, mm 24

Exhaust valve diameter, mm 34

Inlet valve open 13 deg. bTDC

Inlet valve close 30 deg aBDC

Exhaust valve open 20 deg bBDC

Exhaust valve close 14 deg. aTDC

Relative air fuel ratio 1.5

Inlet manifold pressure, atm 1.0

Exhaust manifold pressure, atm 1.0




B40 and B60 was due to complete combustion of oxygenated fuel,

and the lower power output in case of B20 was due to increased

viscosity and presence of relatively less oxygen. However, they

varied the speed from 1200 rpm to 2400 rpm and the power values

were found at a CR of 18. Moreover, the engine design parameters

were also different from the ones that are considered in the present

article. Fig. 1 presents that at CR 14, the peak power occurs at

1400 rpm for all the fuels. The peak power values for diesel, B20,

B40 and B60 are 4.321, 3.954, 4.366, and 4.961 kW respectively.

Fig. 2 shows that at CR 15.5, the peak power occurs for diesel, B20

and B60 at 1400 rpm and for B40, it occurs at 1500 rpm. At CR 17.5

(Fig. 3.), the peak power values occur at 1500 rpm for diesel

(4.69 kW), B20 (4.517 kW), B40 (4.927) with peak brake power

values within the bracket, except for B60, it occurs at 1400 rpm(5.512 kW). From analysis of the model results it was found that the

increased power with B40, B60 is due to increase in the combustion

and expansion works and reduction in the compression work. As

a result, the net work done during the cycle increases and hence,

the power increases. The net work done during the cycle is the sum

of the combustion and expansion work minus the compression

work and the loop work (work done during the gas exchange

process). In the model, early injection was considered for the bio-

diesel blends because the injection timing is advanced due to

higher sound velocity and bulk modulus of biodiesel at low pres-

sure [25]. At 1500 rpm andCR of 17.5,the injectiontiming fordiesel,

B20, B40 and B60 were considered 17, 18, 19 and 20- crank angle

before top dead centre (TDC) respectively. Higher cetane number of

the blends led to the decrease of ignition delay and with slightlymore combustion durations [26]; this ultimately resulted in higher

combustion work in case of the blends (B40 and B60). It can also be

seen from Figs. 4 and 5 that the gas temperature is more during

combustion in case of the blends and hence it results in a higher

pressure during the combustion phase except for the blend B20

that shows slightly less pressure in the later stages of combustion.

This is the reason that combustion works were more in case of the

blends. For B40 and B60, the combustion works were 514.829 and

529.244 J at 1500 rpm and at CR 17.5 as against 505.358 J with

respect to diesel. For B20 it was slightly less than diesel. The

expansion works predicted for diesel, B20, B40, B60 were 381.381,

380.132, 392.606, 415.718 J respectively. Slightly less brake power

in case of the blend B20 was due to reduced combustion and

expansion work.

3.2. Effect of CR on brake power

Figs. 6 and 7 summarize the predicted effect of CR on engine

brake power at two different speeds, viz. 1400 and 1500 rpm. With

increasing CR, the brake power increases for all the fuels. With the

change in CR, engine processes that influence its performance and

efficiency, namely, combustion rate, heat transfer and friction, also

vary. As the CR is increased, the heat loss to the combustion

chamber wall and frictional losses decrease [13]; hence, there is an

improved performance at higher CR. However, there is a limit at

which further increase in CR would not be beneficial as it may lead

to increasing surface to volume ratio and slower combustion;

because at higher CR, the height of the combustion chamber

becomes very small. The brake power results predicted by the

present model also show an increasing trend with CR for all thefuels. At 1400 rpm as shown in Fig. 6, the brake power values for

diesel fuel varied from a minimum of 3.894 kW (at CR 12) to

a maximum of 4.731 kW (at CR 18.5). The corresponding minimum

and maximum values for B20, B40 and B60 are (3.512 kW,

4.599 kW); (4.138 kW, 4.984 kW)and (4.489 kW,5.651 kW)res-

pectively. Fig. 7 presents the brake power values as a function of CR

at 1500 rpm. The minimum and maximum brake power values at

1500 rpm for these fuels i.e. diesel, B20, B40, B60 are (3.794 kW,

4.779 kW); (3.505 kW, 4.631 kW); (3.899 kW, 5.059 kW) and

(4.429 kW, 5.5 kW), respectively. Higher brake power for B40 and

Table 3

Properties of diesel, biodiesel and their blends [11].

Sl. No. Fuel Specific Gravity Kinematic viscosity

(mm2/s)

Calorific value

(MJ/kg)

1 Diesel 0.846 2.6 42.21

2 B20 0.848 3.39 38.28

3 B40 0.856 4.63 37.85

4 B60 0.864 5.42 37.25

2

2.5

3

3.5

4

4.5

5

5.5

600 800 1000 1200 1400 1600 1800 2000

Speed (RPM)

B r a k e P o w e r ( k W )

d iese l B 20 B40 B60

Fig. 1. Brake Power vs Speed at CR 14.

2

2.5

3

3.5

4

4.5

5

5.5

6

600 800 1000 1200 1400 1600 1800 2000

Speed (RPM)


d iese l B20 B40 B60

Fig. 2. Brake Power vs Speed at CR 15.5.

3.00

3.50

4.00

4.50

5.00

5.50

6.00

600 800 1000 1200 1400 1600 1800 2000

Speed (RPM)


d iese l B 20 B40 B60

Fig. 3. Brake Power vs Speed at CR 17.5.




B60 and slightly lower brake power for B20 compared to diesel are

also evident from these figures. This is because these two figures

correspond to a particular constant speed and at constant speed

(1400 rpm and 1500 rpm) the brake power for B40 and B60 is more

and it is slightly less for B20 in comparison to the brake power

obtained for diesel. But for all the fuels the brake power increasedwith CR.

3.3. Effect of speed on brake thermal efficiency

Fig. 8 represents the predicted trend of brake thermal efficiency

as a function of speed at CR 17.5. Brake thermal efficiency decreases

with speed for all the fuels. The blends B20, B40 and B60 present an

increase in brake thermal efficiencycompared to diesel. Since brake

thermal efficiency is the ratio of brake power to the fuel energy

input, therefore, due to increased brake power and less fuel energy

input with biodiesel blends, the brake thermal efficiency is more in

case of the blends. Higher brake thermal efficiency with pure bio-

diesel was reported by Murrilo et al. [27], who compared the

performance of a three-cylinder variable speed (2000–3500 rpm)diesel engine running on blends of diesel and biodiesel derived

from used cooking oil. However, they observed similar behaviour in

case of the blends B10, B30 and B50, but slightly less efficiently in

comparison to diesel, which they attributed to fuel atomization

during injection and its stability during storage, pumping and

injection. Raheman and Phadatare [11] obtained higher brake

thermal efficiency with blends B20 and B40, almost similar

performance with B60 and lower brake thermal efficiency with B80

and B100. Lower calorific value of the blends B60 – B100 together

with increased fuel consumption were reported as reasons for

lower brake thermal efficiency. Although these brake thermal

efficiency results were obtained for karanja methyl ester, but the

engine that was used for evaluating the performance was different

with rated output of 7.5 kW at a speed of 3000 rpm and CR of 16.Moreover, the results were obtained as a function of load at an

average speed of 2525 rpm. In the present work, the increased

brake thermal efficiency in case of the blends (B40 and B60) is due

to comparatively higher brake power and lower calorific value of

these blends. In case of the blend B20, even if the brake power is

slightly less, increased brake thermal efficiency may be due to

calorific value of the blend, which is significantly lower than that of

diesel.

4. Model prediction- analysis and future scope

The calorific value, density, cetane number and the composition

of the fuels are some of the parameters that were taken into

consideration while defining the characteristics of the fuel.However, viscosity is also an important property that plays a major

role particularly in the injection and in the engine combustion

process. The effect of viscosity is best understood if the injection

system and the fuel spray combustion characteristics are consid-

ered in details. Further, It was assumed that the fuel mass is

injected instantaneously and calculation of fuel mass that takes

part in chemical reaction during combustion is based on complete

combustion of the fuel. However, an approach that is more realistic

is the incorporation of a fuel injection rate profile in the model with

a finite injection duration period. The relative air fuel ratio was

500

700

900

1100

1300

1500

1700

1900

2100

150 170 190 210 230

Crank angle (degree)

T e m p e r a t u r e ( K )

diesel B20 B40 B60

Fig. 4. Temperature as a function of crank angle during combustion at 1500 rpm and

CR 17.5.

0

20

40

60

80

150 160 170 180 190 200 210 220 230

Crank angle (degree)

P r e s s u r e ( b a r )

diesel B20 B40 B60

Fig. 5. Pressure as a function of crank angle during combustion at 1500 rpm and CR

17.5.

3.00

3.50

4.00

4.50

5.00

5.50

6.00

10 12 14 16 18 20

CR

b r a k e p o w e r ( k W )

diesel B20 B40 B60

Fig. 6. Brake Power vs CR at 1400 rpm.

3.00

3.50

4.00

4.50

5.00

5.50

6.00

10 12 14 16 18 20

CR


diesel B20 B40 B60

Fig. 7. Brake Power vs CR at 1500 rpm.




considered to be 1.5 for all the fuels at all the speeds under

consideration, although in the real engine situation it will vary with

fuel, speed and other operating parameters. It is observed from the

model results that at CR 17.5 and 1500 rpm, the peak power values

for diesel, B20, B40 and B60 are 4.69 kW, 4.517 kW, 4.927 kW and

5.418 kW respectively. It is generally at this speed, the rated powerof a small conventional diesel engine under consideration is spec-

ified which is of the order of 4.0 kW at CR of 17.5 and at 1500 rpm.

The present model predicts slightly higher brake power in case of

diesel and hence the error is marginal. Considering the fact that the

scope of the present work is to predict the brake power and brake

thermal efficiency trend, it was found that the model has been

successful in predicting the trend correctly for brake power and

brake thermal efficiency at various speeds and CRs for diesel as well

as the blends of diesel and biodiesel.

5. Conclusions

A diesel engine cycle simulation model is developed for pre-

dicting the performance of a single cylinder four stroke dieselengine fuelled by diesel and various blends of diesel and biodiesel.

The brake power and brake thermal efficiency predicted by the

model under varying speed and CR conditions for different fuels are

analyzed and the following conclusions are made based on the

results obtained.

1. Depending upon the engine design parameters, the brake

power of the engine increases with speed, the peak power

occurs at particular speed and any further increase in speed

results in decrease of power. This has been a characteristic of

diesel engine. At CR of 17.5, the engine peak power occurred at

1500 rpm for diesel, B20 and B40; however, for B60 it occurred

at 1400 rpm.

2. The model predicts a higher rate of pressure and temperature

rise for the blends during combustion as compared to diesel.

3. An Increase in brake power was observed in case of the blends

B40 and B60 compared to diesel. However, the model predicts

slightly lower brake power for the blend B20.

4. The blends B20, B40 and B60 present an increase in brake

thermal efficiency compared to diesel. This is due to an increase

in the brake power in case of the blends and their lower calo-

rific values.

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20

25

30

35

40

45

600 800 1000 1200 1400 1600 1800 2000

Speed (RPM)

B r

a k e T h e r m a l E f f i c i e n c y ( % )

diesel B20 B40 B60

Fig. 8. Brake thermal efficiency vs speed at CR 17.5.


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Cycle Simulation