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International Journal of Mechanical Engineering and Technology (IJMET)

Volume 9, Issue 9, September 2018, pp. 1156–1172, Article ID: IJMET_09_09_127

Available online at http://www.iaeme.com/ijmet/issues.asp?JType=IJMET&VType=9&IType=9

ISSN Print: 0976-6340 and ISSN Online: 0976-6359

© IAEME Publication Scopus Indexed

CI ENGINE PERFORMANCE ANALYSIS IN

DUAL FUEL MODE WITH HHO GAS

INDUCTION

P. V. Manu*, S. Jayaraj and A. Ramaraju

Department of Mechanical Engineering,

National Institute of Technology Calicut, Kerala, India

*Corresponding Author

ABSTRACT

High energy content, clean burning, and abundant availability makes hydrogen an

attractive option for using it IC engines as an alternative fuel. Electrolysis is one of

the simplest methods to produce HHO (oxy-hydrogen) gas. A dry cell electrolyzer was

used to generate HHO gas and the generated gas is fed through inlet manifold along

with air in a CI engine. Experiments were conducted on the CI engine with diesel and

diesel + HHO gas as fuels. Engine performance and emission parameters for different

flow rates of HHO gas were considered. It was observed that using HHO as

secondary fuel in CI engine, has positive effects on fuel economy and reduced the

pollutant emissions. A simple two-zone combustion model was developed for the single

cylinder diesel engine considering the effects of premixed and diffusion combustion in

Wiebe’s mass fraction burn profile. Simulations were run for different values of

premixed and diffusion combustion durations and form factors until the simulated

pressure profile matches the experimental profile. Temperatures of burned and

unburned zones were calculated at different loads for straight diesel and dual fuel

operations.

Keywords: Dry cell electrolyzer, HHO gas, Two-zone combustion model.

Cite this Article: P. V. Manu, S. Jayaraj and A. Ramaraju, CI Engine Performance

Analysis in Dual Fuel Mode with HHO Gas Induction, International Journal of

Mechanical Engineering and Technology, 9(9), 2018, pp. 1156–1172.

http://www.iaeme.com/IJMET/issues.asp?JType=IJMET&VType=9&IType=9

1. INTRODUCTION

Energy is a most important factor for the economic development of any country. There is

increasing demand for energy in order to satisfy the enhanced comforts of human life. The

energy sources that we mainly depend are fossil fuels like coal, petroleum, natural gas, etc. A

considerable amount of world energy is utilized by transportation sector. The efficiency of

internal combustion engines has reached its optimum level due to continuous research over

the past many decades. There is no much scope to increase the efficiency of internal

CI Engine Performance Analysis in Dual Fuel Mode with HHO Gas Induction


combustion engines further [1]. Another issue related to the use of fossil fuels is the

environmental pollution caused by the emissions due to the combustion process. The air

pollution caused by fossil fuel combustion affects the environment badly by causing acid rain

and increasing global warming. Air pollution also causes several health hazards including

respiratory issues for humans. The reserves of fossil fuels are decreasing at a greater rate.

Therefore there is a need of alternative fuels which can reduce the energy need and

environmental degradation. Hence, for the growth and development of a country, better

alternative fuels are required. The present energy situation causes active interest in research of

renewable and non-polluting fuels. The term alternative fuel implies that it can be used as an

alternative to conventional fossil fuels. Several attempts were made to find the alternative

fuels which can replace the fossil fuels and doesn’t cause further environmental degradation.

Alternative fuels are available in solid, liquid and gaseous forms. Biomass, biodiesel, LPG,

etc. are some examples of solid, liquid and gaseous alternatives respectively. These fuels can

be directly used in internal combustion engines or as secondary fuels [2, 3]. Some of the

alternative fuels also can generate considerable amount of pollutants.

Among all other alternative fuels, hydrogen is found to be the cleaner fuel [4]. Hydrogen

can be obtained from natural resources like biomass, coal etc. The hydrogen production

techniques available are electrolysis, biomass gasification, thermochemical decomposition of

water and solar photo- electrolysis, etc. [5, 6]. As water resources are abundant, hydrogen

supply also can be considered abundant (as it can be produced from water [7 - 10]). The

combustion of hydrogen produces water vapour only. Hydrogen, as an energy source, has

desirable properties for combustion which includes wider range of flammability limits, higher

flame speed and fast burning velocity (which supports the combustion in lean burn mixtures).

The calorific value of hydrogen is about three times that of the petrol and diesel which

indicates the high level of energy content present in it. As hydrogen requires very low

ignition energy, care should be taken while handling it as fuel. In engine applications, storage

and transportation of hydrogen are the key problems, which need to be addressed [11].

Hydrogen is stored in the compressed form, or using metal hydride technology for an on-

board usage in internal combustion engine, which involves high cost and safety issues.

The carriage of hydrogen as a cryogenic liquid has many challenges in design, safety and

maintenance. On-board generation of hydrogen by water electrolysis will solve the above

issues. The HHO generation process from water is similar to any electrolysis process and

involves an anode, a cathode, DC power source and an electrolyte solution [12, 13]. Hydrogen

and oxygen ions present in water get separated with the application of electric current. When

electric current is applied, hydrogen gas will get generated at the cathode side and oxygen gas

at the anode. Pure water used for electrolysis is not a good conductor of electricity, since it

produces fewer ions. So, in order to improve the conductivity of water in the electrolyzer,

chemicals, such as potassium hydroxide (KOH) is added to water. The KOH splits to produce

K+ ions and OH

- ions in the solution and these ions are responsible to cause conductivity of

the solution [14]. The chemical reactions that happen during water electrolysis are as follows:

Cathode : 2H2O (l) +2e- H2 (g) + 2OH

- (aq)

Anode : 2OH- (aq) ½ O2 (g) +H2O (l) +2e

-

Overall reaction : H2O (l) H2 (g) + ½ O2 (g)

In standard conditions, a theoretical potential difference of 1.23V is needed to perform

electrolysis to split the water and produce the HHO (oxy-hydrogen) gas. This voltage is

corresponding to the Gibbs free energy required for the process at the standard conditions of

298K and 1 bar [15]. Experimental works are aimed at fuel economy and lower pollutant

emissions involve a number of combinations, which are time consuming and cost intensive.

By developing a suitable model for analyzing the engine characteristics, the work may get

P. V. Manu, S. Jayaraj and A. Ramaraju


simplified [16]. In the recent decades, engine combustion models are gaining popularity for

designing better performing engines with lower emissions [17, 18]. In single zone models, the

working fluid in the cylinder is assumed as a thermodynamic system, which undergoes energy

and/or mass exchange with the surroundings. By applying first law of thermodynamics to the

system, the energy released during the combustion process can be obtained [19, 20]. In two-

zone models, the working fluid is imagined to consist of two zones, namely, an unburned

zone and a burned zone. These zones are actually two distinct thermodynamic systems with

energy and mass interactions between themselves and their common surroundings (the

cylinder walls). By applying first law of thermodynamics to the two zones and solving the

resulting simplified equations, we can get the rate of mass fraction burned (or the cylinder

pressure), as a function of the crank angle [21, 22]. Both these models have been used to

predict the in-cylinder pressure as a function of crank angle from an assumed energy release

or mass burned profile (as a function of the crank angle). Another use of these models lies in

determining the energy release/mass burning rate as a function of crank angle from

experimentally obtained in-cylinder pressure data. Multi-zone models take this form of

analysis one step further by considering the energy and mass balances over several zones, thus

obtaining results that are closer to reality.

2. EXPERIMENTAL SETUP

The experimental setup, as shown in Fig. 1, consists of the dry cell electrolyzer, bubbler,

silica gel and DC power supply. Flame arrestor is provided for safety. The generated gas is

fed to the inlet manifold of the engine. The experimental set-up includes a naturally aspirated,

single cylinder, four stroke, water cooled CI engine with data acquisition system for capturing

the in-cylinder pressure for every degree of crank rotation. The values are averaged over 50

cycles and stored in database. The test engine specifications are given in Table 1.

Table 1 Specifications of the test engine.

Manufacturer Kirloskar Engines Ltd.

Type Single cylinder, vertical, four stroke, compression ignition, constant speed,

direct injection, water cooled

Rated power 3.68 kW at 1500 rpm

Bore x stroke 80 mm x 110 mm

BHP 5 hp

Swept volume 553 cc

Compression ratio 16.5:1

The dry cell electrolyzer is provided with electrical energy from a DC power source. The

generated HHO gas was fed to the inlet manifold of diesel engine. By varying the current,

flow rate of HHO can be varied. At different flow rates of HHO and at different loads, various

readings of engine were taken. The engine exhaust emissions were measured using an AVL

make gas analyzer.



Figure 1 Components of experimental setup.

3. EXPERIMENTAL RESULTS

The flow rates of HHO gas at different values of input currents at the electrolyzer were

measured. Experimental and theoretical flow rates were plotted in Fig 2. From this figure, it

can be observed that, the actual flow rate is less than the theoretical flow rate, which can be

attributed to the various resistances within the dry cell electrolyzer.

Figure 2 HHO gas production rate at different current values.

The engine performance and emission parameters for diesel and dual fuel operations were

plotted with respect to the load. From the Fig. 3 and Fig. 4, it can be observed that total fuel

consumption was less and brake thermal efficiency was more for the dual fuel engine. The

reason is the higher flame speed of hydrogen, which helps in complete combustion of diesel.

Figure 3 Variation of diesel consumption at different loads.



Figure 4 Variation of BTE at different loads.

The NOx emission characteristics are shown in Fig. 5. From this figure, it can be observed

that, NOx was decreasing as the HHO flow rate is increasing. Pure water vapor is the product

of hydrogen combustion. It will carry some amount of heat with it. Hence it will decrease the

temperature in the combustion chamber. But after certain flow rate of HHO gas, the energy

release rate due to presence of hydrogen is more due to the higher calorific value. Hence the

NOx emission is expected to increase with the use of HHO gas.

Figure 5 Variation of NOx at different loads.

Figure 6 Graph between load and CO.

Fig. 6 shows that, carbon monoxide is lower for dual fuel operation. Hydrogen has high

flame speed which helps in the complete combustion of diesel, as well. Hence amount of



carbon monoxide in exhaust will decrease as the flow rate of HHO increases.

Correspondingly the amount of carbon dioxide in the dual fuel operation is expected to be

more than that of diesel only operation, which is due to the complete combustion of diesel, as

shown in Fig. 7. From Fig. 8, it can be observed that the amount of oxygen in exhaust gas is

more for dual fuel operation, since HHO gas contains oxygen molecules.

Figure 7 Variation of CO2 at different loads.

Figure 8 Variation of O2 at different loads.

4. THERMODYNAMIC MODELLING

For a simple compressible system at any state, if any two independent intensive properties are

known, then the system is assumed to be completely defined. In such conditions temperature,

pressure and volume are the usual selected parameters as they can be validated by direct

measurement. In the present investigation, out of these three properties volume is considered

as input parameter which is the function of angular displacement of crank. If anyone among

temperature and pressure is simulated then, other can be calculated by the equation of state.

Here the pressure and temperature is determined using the equation of state. From initial value

of pressure, with the help of rate of change of pressure at each point, we can determine

pressure at the subsequent point. Rate of change pressure is a function of the rate of change in

volume and the rate of heat supplied to the system. The following assumptions are involved in

the present modeling:

1. Processes involved in the modelling are quasi-static.



2. Blow by losses are negligible, so that, no exhaust gas leakage is assumed.

3. Working medium behaves similar to an ideal gas at every point of time.

The modeling was done for the engine corresponding to the condition, when all the valves

are closed. It was assumed that, the engine is working on ideal cycle (i.e., inlet valve opens at

TDC and close at BDC, exhaust valve opens at BDC and close at TDC). The compression and

expansion processes were assumed to be polytropic with an index of 1.3.

4.1. Determination of rate of change of pressure

First law of thermodynamics applied for a small change in crank angle dϕ is

Q W dU (1)

For an ideal gas we have,

vdU mc dT (2)

Also, δW = pdV and pV = mRT (3)

Differentiating the ideal gas equation and from Eqn (2), (3), we get,

)( VdppdVR

CdmTCdTmC v

vv (4)

The change in internal energy is,

dmTCVdppdVR

CdU v

v )( (5)

Then the first law equation becomes,

dmTCVdppdVR

CpdVQ v

v )( (6)

Differentiating Eqn (6) with respect to crank angle (ϕ) and by using

CP –CV =R and CP /CV =k (7)

( )

(8)

The rate of change of pressure of the system represented by Eqn (8), is due to the

following:

1. Rate of change of volume dV

d

,

2. Net rate of heat supplied to the system

d

dQ and

3. Rate of change of mass to the system

d

dm

Rate of change of pressure associated with fuel injection is neglected. Hence, the change

in pressure is due to change of volume and net heat supplied to the system.

4.2. Determination of rate of change of volume

The gas volume V in a reciprocating engine can be related to engine geometry as a function of

crank angle. Instantaneous volume of cylinder from the engine kinematics can be written as,



2/122 )sin(cos121

)(

RRV

r

VV dd

(9)

Where, r is the compression ratio, R is ratio of length of connecting rod to crank. The

instantaneous area for heat transfer consists of area of cylinder head, piston, and instantaneous

area of cylinder surface. It is given as

))((4

)(2

salDD

AA ch (10)

Where, ( )s is the instantaneous distance between piston pin and crank pin, given by

21

222 sincos)( alas (11)

4.3. Rate of heat supplied to the system

It consists of rate of heat supplied to the system and rate of heat loss from system to the

cylinder walls. Heat supplied to the system is given as

( )(supplied) *

fd x mdQCV

d d

(12)

Where, ‘x ‘is the mass fraction of fuel burned.

(supplied) * * f

dQ dxCV m

d d

(13)

Where, CV is the calorific value of fuel and dx

d is the rate of change of mass fraction of

the reactants. When rate of change of mass fraction is multiplied with total mass of fuel

supplied in one cycle, it will give the rate of fuel consumption. By multiplying rate of fuel

consumption with calorific value, the rate of heat supplied to the system can be obtained.

The Wiebe’s function [23] for mass fraction burned is given as,

1

*1

mdC t t

x e

(14)

Where, td is the duration of combustion in seconds, t is the time at which burn fraction has

to be determined

ln(1 )dC x (15)

If 0.999dx (considering 99.90 % combustion efficiency), Eqn. (15) gives C = -6.908,

and correspondingly Eqn. (14) becomes

1

6.908*1

m

dt tx e

(16)

The above equation can be written in terms of the burn angle by replacing t

, where, ω

is the constant angular speed

1

ignition6.908*

1

m

dx e

(17)

Now, the rate of mass fraction burned is given by



1

ignition6.908*

ignition6.908( 1)* *

m

d

m

d d

dx me

d

(18)

Where, m is the form factor which determines nature of burning. ignition is the crank angle

at which ignition starts and d is combustion duration.

Variation of rate of mass fraction burned with form factor, which varies from 0.1 to 10 are

shown in Fig 9. For m = 0.1, initially fuel burns faster and then the rate of burning slows

down. Whereas for m = 10, initially rate of combustion is less but at the end it increases. The

actual burning pattern falls in between these two values of m (m = 0.1 to 10). Cumulative

mass fraction burned with respect to the normalized burn angle is shown in Fig. 10. The rate

of heat release follows the same pattern as the mass fraction burned.

Figure 9 Variation of the rate of mass fraction burned at different normalized burn angle.

Figure 10 Variation of the cumulative mass fraction burned at different normalized burn angle.

Diesel combustion consists of both premixed and diffusion phases. Hence, single Wiebe

function does not give accurate results. A double Wiebe function considering premixed and

diffusion combustion can be written as

1

1

908.6

908.6

*)1(908.6

*)1(908.6

dm

d

d

pm

p

p

em

Q

em

Qd

dQ

m

dd

dd

m

pd

p

p

(19)

Where, Qp is the amount of heat release in premixed part of combustion, Qd is the

amount of heat release in diffusion part of combustion, mp is the form factor for premixed



combustion, md is the form factor for diffusion combustion, p is the duration of premixed

combustion, d is the duration of diffusion combustion.

4.4. Wall heat transfer

Woschini [24] gave a set of empirical relations that predicts the heat transfer coefficient

between system and the cylinder walls. According to Newton’s law of cooling the wall heat

transfer can be written as

( )wallQ hA T (20)

The heat transfer coefficient given by Woschini is

0.2 0.8 0.55 0.83.26* * * *h D p T u (21)

Where, the parameter is given by the equation

1 2 ( )r

m

r r

VTu c u c p p

p V

(22)

In the Eqn. (22), ̅ is mean piston speed and Tr, pr and Vr are reference temperature,

pressure and volume (respectively). For the compression and expansion processes, Watson

and Janota [24] suggested modeling the motored cylinder pressure as a polytropic process.

For compression, 1 2.28c and 2 0c

For expansion, 1 2.28c and

3

2 3.24 10c

4.5. Ignition delay

Ignition delay in diesel engine is defined as the time lag between start of injection and start of

combustion. It depends on temperature and pressures at start of injection, type of fuel, load

etc. Wolfer [3] gave an empirical correlation for ignition delay in terms of the above

parameters.

0.7352exp 0.0187 8.551 * exp(4626.44 / )ms i LCN p T (23)

1 {0.008* * } *L iT f L T (24)

Where, ms is ignition delay in milliseconds, CN is cetane number of fuel, ip pressure in

the cylinder at the start of injection and LT is temperature of gas in the cylinder at the start of

injection at a particular load. The parameter f is 0.5 for stationary engine. So ignition delay in

terms of crank angle can be written as

( ) ( )*0.006*CA ms N (25)

4.6. Two zone combustion model

Single zone combustion models are simplest but inadequate in estimating the accurate

estimation of pollutant formation. In single zone models the temperature at each time step is

the average temperature in the cylinder. But the rate of chemical reaction depends on the

temperature at that particular location. Hence the two zone models are becoming popular in

which the system is divided into two zones (called burning zone and unburned zone). Pressure

at every crank angle is determined from the single zone model and that is used to estimate the



temperatures of burned and unburned zones. Temperatures of burned zones are used to

estimate the pollutants. The assumptions included in developing the two zone combustion

model are:

i. The burned and unburned zones are ideal gases with different properties,

ii. No heat transfer occurs from the burned to the unburned zone and vice versa,

iii. Enthalpy associated with injected fuel is usually not significant and hence ignored,

iv. Instantaneous pressure in both the zones is the same and

v. The work required to transfer fluid from the unburned zone to the burned zone is

negligible.

The mass of burned and unburned zones at ith

iteration in computation are given as

( ) ( 1) bb b c

dXm i m i i m

d

(26)

( ) ( 1) uu u c

dXm i m i i m

d

(27)

Law of conservation of mass is written as

b u cm i m i m (28)

1

( ) ( 1) ( )( )

( 1) ( 1)

nu u

u

u

m i V i P iV i

m i P i

(29)

( ) ( ) ( )u bV i V i V i (30)

By solving the above equations for every crank angle and by using ideal gas equation, we

can write temperatures of burned zone and unburned zones as

( ) ( )( )

( )

bb

b

P i V iT i

m i R

and

( ) ( )( )

( )

uu

u

P i V iT i

m i R

(31)

5. SIMULATION RESULTS

Ignition delay was calculated using the Wolfer correlation and compared with the

experimental results as shown in Fig. 11. The correlation predicts delay period with

reasonable accuracy. The heat release rates were calculated from double Wiebe function at

different values of combustion durations (Δθc) and form factors (mp and md). The following

Figs. 12 and 13, are the heat release rate graphs, plotted for different values of diffusion form

factors md at 100 premixed and 50

0 diffusion combustion durations and mp = 0.1.



Figure 11 Variation of the ignition delay with respect to load.

Figure 12 Variation of heat release rate with respect to crank angle (mp = md = 0.1).

Figure 13 Variation of heat release rate with respect to crank angle (mp = 0.1, md = 0.5).

Attempts were made to determine the combustion duration and values of form factors for

premixed and diffusion combustion phases. The Figs. 14, 15 and 16, shows the pressure traces

for a combustion duration of 400

CA (100 premixed and 30

0 diffusion), for mp = 0.1, 0.5 and 5

and at different values of md. From these graphs, it can be concluded that no curve is matching

the experimental profile. Hence the values of combustion durations and form factors (mp and

md) were varied to obtain the trend which matches with the experimental profile.



Figure 14 Variation of pressure with respect to crank angle (Δθc = 400 CA, mp=0.1).

Figure 15 Variation of pressure with respect to crank angle (Δθc =400 CA, mp=0.5).

Figure 16 Variation of pressure with respect to crank angle (Δθc =400 CA, mp=5).

Similar attempts made for a combustion duration of 600 CA (10

0 premixed and 50

0

diffusion), for mp = 0.5, 5 and at different values of md are shown in Figs. 17 and 18.



Figure 17 Variation of pressure with respect to crank angle (Δθc =600 CA, mp=0.5).

Figure 18 Variation of pressure with respect to crank angle (Δθc =600 CA, mp=5).

A number of combinations of premixed and diffusion combustion durations and form

factors (mp and md) were checked. Out of all these combinations, combustion duration of 600

CA (100 premixed and 50

0 diffusion), for mp = 5 and md = 0.5, gave the p–ϕ trend close to

experimental one, which is shown in the Fig. 19. Variation of mass burned and unburned

zones with respect to crank angle for a combustion duration of 600 CA (10

0 premixed and 50

0

diffusion), for mp = 5 and md = 0.5, is shown in Fig. 20. Initially mass of burned zone is zero

and unburned zone is equal to the mass of gases in the cylinder. As combustion progresses,

mass of unburned zone decreases and mass of burned zone increases. Total mass of gases in

the cylinder remain constant.

Figure 19 Graph between crank angle and pressure.



Figure 20 Variation of unburned and burned gases mass fraction.

Figure 21, shows variation of burned and unburned zone temperatures with respect to

crank angle at full load. Temperature of burned zone can be used for prediction of emissions.

The temperatures of burned and unburned zones for dual fuel operation at different flow rates

of HHO are shown in the Figs. 22 and 23. As the HHO flow rate increases, the temperature of

the burned zone increases. Increase in HHO flow rate increases the peak temperature in the

cylinder, which is due to the fact that hydrogen is possessing higher calorific value.

Figure 21 Variation of temperature for diesel operation at full load condition.

Figure 22 Variation of temperature for dual fuel operation (0.14 LPM of HHO) at full load condition.



Figure 23 Variation of temperature for dual fuel operation (0.18 LPM of HHO) at full load condition.

6. CONCLUSIONS

As an automobile fuel, hydrogen is getting importance because of its clean combustion

characteristics. From experiments, it is confirmed that diesel consumption is decreasing and

brake thermal efficiency is increasing for dual fuel operation. The exhaust emission of CO,

NOx and smoke are less for dual fuel operation. Whereas CO2 in exhaust is found to be more

for HHO operation because of complete combustion of diesel. A simple two zone combustion

model was developed and temperatures of burned and unburned zones were calculated.

Combustion duration of 600 CA, mp = 5 and md = 0.5 gave the better match with the

experimental pressure trace. The temperatures of burned zone can be used for prediction of

emissions. HHO as a secondary fuel in the diesel engine has positive effects on engine

performance and exhaust emissions. In the near future, HHO gas has the capability to replace,

at least partially, the use of fossil fuels.

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