performance evaluation of low heat rejection diesel engine ...performance of the engine with superni...

INTERNATIONAL JOURNAL OF APPLIED ENGINEERING RESEARCH, DINDIGUL Volume 1, No 3, 2010

© Copyright 2010 All rights reserved Integrated Publishing Association

REVIEW ARTICLE ISSN 09764259

428

Performance Evaluation of Low Heat Rejection Diesel Engine with Pure Diesel

Murthy P.V.K 1 , Murali Krishna M.V.S. 2 , Sitarama Raju A 3 , Vara Prasad C.M. 4 Srinivasulu N.V. 2

1 Vivekananda Institute of Science and Information Technology, Shadnagar, Mahabubnagar

2 Mechanical Engineering Department, Chaitanya Bharathi Institute of Technology, Gandipet, Hyderabad.

3 Mechanical Engineering Department, J.N.T. University, Kukatpally, Hyderabad

4 Sreenidhi Institute of Science and Technology, Na Yampet, Ghatkeswar Mandal, Hyderabad

[email protected]

ABSTRACT

Investigations are carried out to evaluate the performance of a low heat rejection (LHR) diesel engine consisting of air gap insulated piston with 3mm air gap, with superni (an alloy of nickel) crown, air gap insulated liner with superni insert and ceramic coated cylinder head with pure diesel operation with varied injection timing and injection pressure. Performance parameters are determined at various magnitudes of brake mean effective pressure. Pollution levels of smoke and oxides of nitrogen (NOx) are recorded at the peak load operation of the engine. Combustion characteristics of the engine are measured with TDC (top dead centre) encoder, pressure transducer, console and special pressurecrank angle software package. Zero dimensional, multizone combustion model is assumed to predict combustion characteristics and validated with experimental results. LHR engine showed deteriorated performance at recommended injection timing and pressure and improved performance at advanced injection timing and higher injection pressure, when compared with conventional engine (CE). At peak load operation, brake specific fuel consumption (BSFC) decreased by 12%, while smoke levels by 16% and NOx levels increased by 34% with LHR engine at an injection timing of 32 o bTDC (before top dead centre) and an injection pressure of 270 bars, in comparison with CE operating at an injection timing of 27 o bTDC, and an injection pressure

of 190 bars.

Keywords: Low heat rejection, Performance, Pollution levels, Combustion characteristics, Zero dimensional multizone combustion model.

NOMENCLATURE

a Constant used in Annand’s equation A Total heat transfer surface area in m 2 b Constant used in Annand’s equation BDC Bottom Dead Centre BMEP Brake mean effective pressure in bar BSFC Brake specific fuel consumption in kg/hkW bTDC Before top dead centre




429

BTE Brake thermal efficiency in percentage c Constant used in Annand’s equation CE Conventional engine CV Calorific value of the fuel in kJ/kg. Dl Cylinder bore in metre dn Diameter of nozzle in metre Dsm Sauter mean diameter of fuel particle in micrometers EGT Exhaust gas temperature in o C HSU Hartridge smoke units IVC Inlet valve closing J Total number of radial divisions K Thermal conductivity in W/mK L Latent heat of evaporation of fuel in J/kg, LHR Low heat rejection M Mass MRPR Maximum rate of pressure rise in bar/degree N / Number of droplets NOx Oxides of nitrogen O/F Oxygen to fuel ratio O2 Oxygen P Pressure in mega Pascal P (θ) Instantaneous gas pressure in bar Phi Equivalence fuelair ratio PP Peak pressure in bar Q Heat transfer in W

1 inj Q Amount of fuel delivered per cycle per second in m 3 / sec

R Gas constant in kJ /kgK R Universal gas constant kJ/ K, r Radial coordinate of cylindrical coordinate system Re Reynold’s number rs Radius of droplet in micrometers, SMD Sauter mean diameter of droplet in microns T Surrounding temperature in Kelvin tbu Brake up time for element in sec TDC Top Dead Centre Tl Temperature of liquid fuel in Kelvin. TOMRPR Time of occurrence of maximum rate of pressure rise in degrees

TOPP Time of occurrence of peak pressure in degrees

TR Theoretical Result through Computer Prediction Tw Wall temperature in Kelvin

t Ignition delay of gaseous mixture

V Volume or domain V ‘ Velocity in m/s V 1 inj Velocity of injection




430

VE Volumetric efficiency in percentage Vdis Displacement volume in m 3 W Molecular weight Xij Spray tip penetration in axial direction in meters Y Mass fraction of chemical species 27 o bTDC Manufacturer’s recommended injection timing 190 bars Manufacturer’s recommended injection pressure

Subscripts

a air ae Air entrainment b burnt bu brake up c combustion dis Displacement e entrainment f Fuel g gas fg Gaseous fuel h cylinder head i index for an element in the axial direction inj injection j index for an element in the radial direction l liquid L Latent heat of vaporization li Liner NS Index for chemical species vary from 1 to 5 referring to fuel

(Referring to the fuel, O2, N2, CO2 and H2O respectively) n net o reference state p Piston St Stoichiometric ratio s surface of droplet u unburnt v vapor w wall 298 Datum temperature in Kelvin

Greek symbols θ Crank angle in degrees δ Differential µ Dynamic viscosity in Ns/m 2

∆ Corresponding to change in a quantity θ i Crank angle corresponding to the location of the package θ o Crank angle corresponding to one complete cycle in degrees




431

β′ Spalding number ρa Density of air in kg/m 3

ρg Gas density in kg/m 3

ρl Liquid density in kg/m 3

ρs Density of droplet in kg/m 3

τ Ignition delay in milliseconds

1. Introduction

In the scenario of increase of vehicle population at an alarming rate due to advancement of civilization, use of diesel fuel in not only transport sector but also in agriculture sector leading to fast depletion of diesel fuels and increase of pollution levels with these fuels, efficient fuel utilization has become pertinent for the engine manufacturers, users and researchers involved in the combustion research. While search for alternate fuels is continuing, researchers are also attempting to find different techniques of efficient fuel utilization in diesel engines.

It is well known fact that about 30% of the energy supplied is lost through the coolant and the 30% is wasted through friction and other losses, thus leaving only 30% of energy utilization for useful purposes. In view of the above, the major thrust in engine research during the last one or two decades has been on development of LHR engines. Several methods adopted for achieving LHR to the coolant are i) using ceramic coatings on piston, liner and cylinder head ii) creating air gap in the piston and other components with lowthermal conductivity materials like superni, cast iron and mild steel etc. Krishnan e al. and Wade et al. observed in improvement in thermal efficiency with ceramiccoated component with diesel as fuel. Woschni et al. and Cole et al. reported reduction in BSFC with an air gap insulated piston at part loads. Creating an air gap in the piston involved the complications of joining two different metals. Though Parker et al. observed effective insulation provided by an air gap, the bolted design employed by them could not provide complete sealing of air in the air gap. Rama Mohan made a successful attempt of screwing the crown made of low thermal conductivity material, nimonic (an alloy of nickel) to the body of the piston, by keeping a gasket, made of nimonic, in between these two parts.

The concept of LHR engine is to reduce heat loss to coolant and that too specifically from the piston top to the body of the piston. It should be expected that the thickness of the air gap play an important role on the insulation effect in LHR engines. Rama Mohan and Dhinagar et al. worked in this direction in order to get higher insulation effect. In order to increase the degree of the insulation, air gap is not only created in the piston but also in the liner. Murali Krishna conducted experiments on LHR engine which consisted of air gap insulated piston with superni crown and air gap insulated liner with superni insert with advanced injection timings and increased injection pressure with different alternate fuels like alcohols and non edible vegetable oil and reported improved performance with LHR engine. While Seigla et al. and Miyairi et al. found deterioration of the engine performance with retarded injection timings, Dhinagar et al. claimed improved performance with retarded injection timing in LHR engine with diesel as fuel.




432

The complexities involved like heterogeneous combustion of diesel fuel and air in minute fraction of time, various sizes of fuel droplets and their penetration and evaporation etc., make diesel combustion usually not amenable for effective modelling. Miyairi used a two zone combustion model and a computer simulation model was used by Rafiqul Islam et al for predicting the performance of the ceramiccoated direct injection (DI) diesel engines. Ramamohan et al and Murali Krishna employed zero dimensional multizone combustion models for predicting the performance of the LHR engine and validated the theoretical results with experimental data and found 57% deviation between these two.

The present paper attempts to evaluate the performance of LHR engine, which contains air gap insulated piston, air gap insulated liner and ceramic coated cylinder head with varying engine parameters of change of injection pressure and injection timing and compared with CE at recommended injection timing and injection pressure. An attempt is made to develop a zero dimensional, multizone combustion model and correlate theoretical analysis with experimental results.

2. Experimental Programme

Figure1gives the details of insulated piston, insulated liner and ceramic coated cylinder head employed in the experimentation. LHR diesel engine contains a twopart piston; the top crown made of low thermal conductivity material, superni90 screwed to aluminum body of the piston, providing a 3mmair gap in between the crown and the body of the piston. The optimum thickness of air gap in the air gap piston is found to be 3mm [6], for better performance of the engine with superni inserts with diesel as fuel. A superni90 insert is screwed to the top portion of the liner in such a manner that an air gap of 3mm is maintained between the insert and the liner body. At 500 o C the thermal conductivity of superni90 and air are 20.92 and 0.057 W/mK respectively. Partially stabilized zirconium (PSZ) of thickness 500 microns is coated by means of plasma coating technique. Experimental setup used for the investigations of LHR diesel engine with pure diesel is shown in Fig.2. CE has an aluminum alloy piston with a bore of 80mm and a stroke of 110mm. The rated output of the engine is 3.68 kW at a rate speed of 1500 rpm. The compression ratio is 16:1 and manufacturer’s recommended injection timing and injection pressures are 27 o bTDC and 190 bar respectively.

The fuel injector has 3 holes of size 0.25mm. The combustion chamber consists of a direct injection type with no special arrangement for swirling motion of air. The engine is connected to electric dynamometer for measuring brake power of the engine. Burette method is used for finding fuel consumption of the engine. Airconsumption of the engine is measured by airbox method. The naturally aspirated engine is provided with watercooling system in which inlet temperature of water is maintained at 60 o C by adjusting the water flow rate. The engine oil is provided with a pressure feed system and no temperature control is incorporated, for measuring the lube oil temperature. Copper shims of suitable size are provided in between the pump body and the engine frame, to vary the injection timing and its effect on the performance of the engine is studied, along with the change of injection pressures from 190 bar to 270 bar (in steps of 40 bar) using nozzle testing device.




433

1. Crown 7 Insert 2. Gasket 8. Air gap 3. Air gap 9. Liner 4. Body 5. Ceramic coating 6. Cylinder head

Insulated piston Insulated liner Ceramic coated cylinder head

Figure 1: Assembly details of insulated piston, insulated liner and ceramic coated cylinder head

1.Engine, 2.Electical Dynamo meter, 3.Load Box, 4.Orifice meter, 5.Utube water manometer, 6.Air box, 7.Fuel tank, 8, Three way valve, 9.Burette, 10. Exhaust gas temperature indicator, 11.AVL Smoke meter, 12.Netel Chromatograph NOx Analyzer, 13.Outlet jacket water temperature indicator, 14. Outletjacket water flow meter, 15.Piezoelectric pressure transducer, 16.Console, 17.TDC encoder, 18.Pentium Personal Computer and 19. Printer.

Figure 2: Experimental Setup




434

The maximum injection pressure is restricted to 270 bar due to practical difficulties involved. Exhaust gas temperature (EGT) is measured with thermocouples made of iron and iron constantan. Pollution levels of smoke and NOx are recorded by AVL smoke meter and Netel Chromatograph NOx analyzer respectively at the peak load operation of the engine. Piezo electric transducer, fitted on the cylinder head to measure pressure in the combustion chamber is connected to a console, which in turn is connected to Pentium personal computer. TDC encoder provided at the extended shaft of the dynamometer is connected to the console to measure the crank angle of the engine. A special Pθ software package evaluates the combustion characteristics such as peak pressure (PP), time of occurrence of peak pressure (TOPP), maximum rate of pressure rise (MRPR) and time of occurrence of maximum rate of pressure rise (TOMRPR) from the signals of pressure and crank angle at the peak load operation of the engine. Pressurecrank angle diagram is obtained on the screen of the personal computer

3. Combustion Modelling

A zerodimensional, multizone model is attempted to predict the performance of LHR diesel engine, with air gap insulated piston and liner. However, there are certain assumptions suck as i) There is no interaction between two elements, ii)Pressure is uniform over the entire combustion chamber, iii) Fuel jet breaks into droplets right at the exit plane of the nozzle and iv) Injection pressure and injection rate are constant over a cycle. The concept of dividing spray is similar to that of Hiroyasu [15].The model is closed cycle simulation and has been divided into five submodels, (1) fuel injection (2) spray penetration and air entrainment (3) spray evaporation (4) combustion model (5) heat transfer model. The spray emerging from the injection nozzle in the form of a cone is divided into number of small elements which are specified by two variables i and j in axial and radial directions. The air package contains fresh air and residual gas for diesel operation is identified by a single index i=1. The number of droplets and their Sauter mean diameter in each element are computed in the following manner [15] in fuel injection sub model.

( ) ( ) 1 inj

0.12 a

0.135 inj j i, SM Q ρ P) (23.9)(P D − − = (1)

( ) ∫ +

= ′ 1 i

i

t

t 3 j i, SM

1 inj

D 6 π J

dt Q N (2)

The location of each element in space is determined from the following relation [15]

The brake up time for element can be found from the following relation.

( ) 0.5 g bu

ΔP ρ

d ρ 29 t l = (3)

Where tθI <tbu , the spray tip penetration in axial direction (Xi,j) direction is

given by [15]




435

1) (j for t ρ P 2 0.39 X θi

5 . 0

j i, =

∆ =

l (4)

Where tθi > tbu, the spraying penetration in axial direction is given by

( ) 2 / 1 θi n

4 / 1

g j i, t d

ρ ΔP 2.95 X

= for (j=1)(5)

The radial location of other element is given by following relation [15],

( ) 2 3 1) (j 8.557x10 Exp j i, i X − − −

= for ( j > 1) (6)

Taking the equation 4, and differentiate w.r.t. time,

θi

j i, j i,

t X

dt dX

= (7)

Taking the equation 5, and differentiate with respect to t time,

for tθi > tbu

θi

j i, j i,

2t X

dt dX

= (8)

The air entrainment into each package is obtained by momentum balance.

The conservation of momentum for each package can be expressed as

[ ] dt

dX M M V M j i,

ae f 1 inj f j i, j i, j i,

+ = (9)

Simplifying the above equation,

dt dX

M dt

dX M V M j i,

ae j i,

f 1 inj f j i, j i, j i,

+ =

Hence entrainment of air in an element is given by




436

− = 1

dt dX V

M M j i,

1 inj

f ae j i, j i, (10)

The rate of air entrainment j i, ae M & is obtained by differentiating the above equation with

respect to time,

By assuming constant injection velocity and injection rate, as the rate of fuel injected/ degree of crank rotation is the function of injection cam velocity, diameter of injector plunger and flow area of tip orifices,

Therefore, 0 M 0; V j i, f

1 inj = = & &

[ ] j i, j i, j i, f ae 1

j i,

1 j i,

ae M M V V

M + − = &

& (11)

The entrainment of individual species into each package can be found from the mass

fraction of that species in the air package (i=1)

=

dt

dM Y

dt

dM j i,

j i, j i, ae

NS NS

(12)

The evaporation rates of droplet in an element are computed using Spalding’s droplet

evaporation relation [16].

[ ] ) β log(1 D ρ r 4 dt dM 1

s s s f ′ + = π (13)

The rate of decrease droplet radius is given by following expression [16]

( ) [ ] β 1 log r D

ρ ρ

dt dr

s

1 s

l

s s ′ + = (14)

Total evaporation rate in element is given by

dt

dM N

dt

dM j i, j i, f

j i, v

′ = (15)




437

The rate of heat removal from gas phase due to latent heat of evaporation is given by

dt

dM L

dt

dQ j i, j i, v v

= (16)

Where L is obtained from the following expression [17]

L = 9764.2 [1799.4 – 1.8Tl] 0.5 (17)

The mass of air and mass of gaseous fuel determines the equivalence ratio (Phi) in the

gaseous phase in the element as

)st /M (M M M

Phi a fg

a fg =

In an element after short period of time from fuel injection, ignition occurs in the gaseous mixture. The ignition delay of gaseous mixture at varying temperature, pressure and equivalence ratio is obtained from

t = 4 x 10 3 (P) 2.5 (Phi) –1.04 Exp(4000)/T (18)

When once the delay period is over, depending upon the availability of oxidant and

vaporized fuel, the chemical reaction is assumed to take place at stoichiometric proportion.

Heat release rate due to combustion of fuel is shown by

t Cv M

dt dQ f c = for Phi < 1

) F O ( t .Cv M

dt dQ 02 c = for Phi > 1

Heat transfer to cylinder walls is obtained by using Annand’s relation [18]. This relation

considers the net heat transfer as the summation of both convective and radiative heat transfer.

[ ][ ] ( ) 4 w

4 w

b w T T Ac T T (Re) D aA.K

dt dQ

− + − = l

(19)

A = Total surface area exposed to heat transfer

= Ap + Ali + Ah




438

The wall temperature is obtained based on FEM analysis carried out for LHR engine.

Global Equations

The net heat release in the element is given by

dt

d

dt

dQ

dt

d

dt dQ j i, j i, j i, w v C n

Q Q − − = (21)

The temperature of an element at any instant is obtained by applying energy equation to each

element.

R) M(C R MT h) (h M h) (h M p v Q T

v

f v a ae n

+ − − + − + +

= & & & & &

& (21)

For air package

w n a v ae Q Q 0; R ; h h 0; M ; M M & & & & & & = = = = − = (22)

By substituting the conditions for air package given in equation (21)

R) M(C P V Q

T v

w

+ +

= & &

& (23)

Pressure is obtained from the conservation of momentum

∑ ∑ +

=

j i, j i, j i, j i, j i, j i, T R M T R M [

V V 1 p

l

& & &

)] V V P( M.C

R T M ) h (h M ) h (h M p V Q MR

j i,

j i, j i, j i, j i, j i,

p

f v a ae n

l & &

& & &

− −

− − + − + +

+ ∑

• •

(24)

Rate of change of gas constant is computed by following relation

∑ = NS

NS

W Y

R R (25)

By differentiating the equation 25,




439

∑

− =

M M Y

M M

W 1

dt dR ns ns

ns

& & R (26)

By knowing pressure and temperature of particular element, its volume can be found by applying the equation of state to individual element.

p T R M

V j i, j i, j i, j i, = (27)

After finding the volume of all elements including the air element, the volume constraint

should be checked. The volume constraint (where the liquid volume is neglected) is given by

∑ = cyel j i, V V (28)

Equations 12,13,14,23 and 24 formed the set of governing equations for pure diesel or non edible version oil operations formed the set of governing equations which are solved using RungeKutta fourth order scheme. Computerization started at IVC and closed at EVO. Up to the point of the injection the process is pure compression and the main computations described in the model are computed after this instant (point of injection).

Results and Discussion

A.Performance Parameters

The variation of brake thermal efficiency (BTE) with brake mean effective pressure (BMEP) in the conventional engine (CE) with pure diesel, at various injection timings at an injection pressure of 190 bar, is shown in Fig.3. BTE increased with the advancing the injection timings in CE for entire range of the load, due to early initiation of combustion as delay period increased with advancing of the injection timing. The optimum injection timing is obtained by based on higher thermal efficiency. The maximum BTE is observed when the injection timing is advanced to 33 o bTDC in CE. Panchapakeshan et al. [19] and Rajan et al. [20] also observed the same trend. The variation of BTE with BMEP in the LHR engine with pure diesel at various injection timings at an injection pressure of 190 bar, is shown in Fig.4. BTE decreased at all loads in LHR engine at the recommended injection timing when compared with CE. As the combustion chamber is insulated to greater extent, it is expected that high combustion temperatures would be prevalent in LHR engine. It tends to decrease the ignition delay thereby reducing premixed combustion as a result of which, less time is available for proper mixing of air and fuel in the combustion chamber leading to incomplete combustion, with which BTE decreased at all loads. More over at this load, friction and increased diffusion combustion resulted from reduced ignition delay. Increased radiation losses might have also contributed to the deterioration. Higher magnitude of BTE at all loads including 100% full load is observed when the injection timing is advanced to




440

32 o bTDC in LHR engine. Further advancing of the injection timing resulted in increase in fuel consumption due to longer ignition delay. Hence it is concluded that the optimized performance of LHR engine is achieved at an injection timing of 32 o bTDC.

0

5

10

15

20

25

30

35

0 2 4 6

BTE,%

BMEP,bar

CE‐Injection Timing‐ 27bTDC




Figure 3: Variation of brake thermal efficiency (BTE) with brake mean effective pressure (BMEP) in conventional engine (CE) at different injection timings

0

5

10

15

20

25

30

35

0 2 4 6

BTE,%

BMEP,bar


LHR‐InjectionTiming‐ 27bTDC

LHR‐Injection Timing‐ 29bTDC



Figure 4: Variation of brake thermal efficiency (BTE) with brake mean effective pressure (BMEP) in low heat rejection (LHR) engine at different injection timings




441

0

100

200

300

400

500

600

0 2 4 6

EGT,Deg

C

BMEP,bar

1‐CE‐Injection Timing‐ 27bTDC

2‐LHR‐Injection Timing‐27bTDC



Figure 5: Variation of exhaust gas temperature (EGT) with brake mean effective pressure (BMEP) in conventional engine (CE) and low heat rejection (LHR) engine at recommend injection timing and optimized injection timings

0 0.5 1

1.5 2

2.5 3

3.5 4

4.5 5

0 2 4 6

CL,k W

BMEP,bar



3‐CE‐InjectionTiming‐ 33bTDC


Figure 6: Variation of coolant load (CL) with brake mean effective pressure (BMEP) in conventional engine (CE) and low heat rejection (LHR) engine at recommend injection timing and optimized injection timings

Higher fuel injection pressures increase the degree of atomization. Performance of the engine is evaluated with varying injection pressure from 190 to 270 bars and injection timing is advanced from 27 to 34 o bTDC for CE and LHR engine.




442

70

75

80

85

90

95

0 2 4 6

VE,%

BMEP,bar





Figure 7: Variation of volumetric efficiency (VE) with brake mean effective pressure (BMEP) in conventional engine (CE) and low heat rejection (LHR) engine at recommend

injection timing and optimized injection timings

0

10

20

30

40

50

60

70

0 2 4 6

Smok

e,HSU

BMEP,bar



3‐CE‐injection Timing‐ 33bTDC


Figure 8: Variation of smoke intensity in Hartridge Smoke Unit (HSU) with brake mean effective pressure (BMEP) in conventional engine (CE) and low heat rejection (LHR)

engine at recommend injection timing and optimized injection timings




443

0

200

400

600

800

1000

1200

1400

1600

0 2 4 6

Nox

,PPM

BMEP,bar





Figure 9: Variation of NOx levels with brake mean effective pressure (BMEP) in conventional engine (CE) and low heat rejection (LHR) engine at recommend injection

timing and optimized injection timings

The improvement in the BSFC at peak load at higher injection pressure is due to improved fuel spray characteristics. However, the optimum injection timing is not varied even at higher injection pressure with LHR engine, unlike CE. Hence it is concluded that the optimum injection timing is 33 o bTDC at 190 bar, 32 o bTDC at 230 bar and 31 o bTDC at 270 bar for CE. The optimum injection timing for LHR engine is 32 o bTDC irrespective of injection pressure. Fig.5 shows the variation of the exhaust gas temperature (EGT) with BMEP in the CE and LHR engine with diesel operation at the recommended and optimized injection timings at an injection pressure of 190 bar. LHR engine, at the recommended injection timing recorded higher EGT at all loads compared with CE at the recommended injection timing.

This indicated that heat rejection is restricted through the piston, liner and ceramic coated cylinder head, thus maintaining the hot combustion chamber as result of which the exhaust gas temperature increased. Exhaust gas temperatures decreased when the injection timing is advanced to 32 o bTDC at all loads (except at peak load) with LHR engine, when compared with CE at the recommended injection timing. LHR engine, at the optimum injection timing recorded marginally higher EGT (4.8% higher) at the full load compared with CE at the recommended injection timing. The magnitude of EGT decreased at all loads in CE at its optimum injection timings, compared with the same version of the engine at the recommended injection timing.

This is because, when the injection timing is advanced, the work transfer from the piston to the gases in the cylinder at the end of the compression stroke is too large, leading to reduce in the magnitude of EGT. The variation in the magnitude of coolant load with BMEP in CE and LHR engine with pure diesel, at the recommended and optimum injection timings at an injection pressure of 190 bar, is shown in Fig.6. Coolant load increased with the increase of load in CE and LHR engine. LHR engine gave lesser coolant load upto 80% of the peak load,




444

when compared with convectional engine. Air being a bad conductor offers thermal resistance for heat flow through the piston and liner.

Conv. Predicted Conv. Experimental LHR EnginePredicted

LHR EngineExperimental Injection Pressure190 bars

Injection Timing27 deg.bTDC

20

0

20

40

60

80

120 240 360 480 600

Crank Angle Deg.

Pres

sure

, bar

s

Figure 10: Comparison of typical pressurecrank angle data predicted from the multizone combustion model and the experimental values obtained from the special Pθ

software package and TDC encoder for the case of conventional and LHR engines at the recommended injection timing and at the recommended injection pressure

It is therefore evident that thermal barrier provided in the piston and liner resulted in reduction of coolant load upto 80% of the full load. Beyond 80% of the full load, coolant load in LHR engine increased over and above that of the CE, with which efficiency is deteriorated at peak load of LHR engine, when compared with CE. This is because in cylinder, the heat rejection at full load is primarily due to unburnt fuel concentration near the combustion chamber walls. The airfuel ratio got reduced to a reasonably low value at this load confirming the above trend. However, when heat rejection calculations of coolant are made, the heat lost to lubricant should also be considered. As in the present investigations the lubricant heat loss is not considered, this aspect is not depicted in coolant load calculations. Wallace et al. and Rama Mohan , Murali Krishna also observed the same trend at full load operation. Coolant load decreased in LHR engine at its optimum injection timing at all loads when compared with both versions of the engine at the recommend injection timing. This is due to decrease of combustion temperatures in LHR engine. In case of CE, unburnt fuel concentration reduced with effective utilization of energy, released from the combustion, coolant load increased marginally at all loads due to increase of gas temperatures, when the injection timing is advanced to the optimum value. Decrease of gas temperatures in the LHR engine with the increase of injection pressure any way decreased coolant load and exhaust




445

gas temperatures. Coolant load increased marginally in CE, while it decreased in LHR engine with the advancing of the injection timing and increasing of the injection pressure. This is due to the fact with increase of injection pressure with CE, increased nominal fuel spray velocity resulting in better fuelair mixing with which gas temperatures increased. Chandorkar et al [21] confirmed that increase in injection pressure increased the cooling rate in CE. The reduction of coolant load in LHR engine is not only due to the provision of the insulation but also it is due to better fuel spray characteristics and increase of airfuel ratios causing decrease of gas temperatures and hence the coolant load.

Figure 11: Variation of airzone temperature and gaszone temperature with respect to crank angle in CE and LHR at recommended and optimized injection timings

The variation in the magnitude of volumetric efficiency (VE) with BMEP in CE and LHR engine with pure diesel, at recommend and optimum injection timings an injection pressure of 190 bars, is shown in Fig.7. VE decreased with the increase of BMEP in both versions of the engine. This is due to increase of gas temperature with the load. At the recommended injection timing, VE in LHR engine decreased at all loads when compared with CE. This is due increase of temperature of incoming charge in the hot environment created with the provision of insulation, causing reduction in the density and hence the quantity of air. However, this variation in VE is very small between these two versions of the engine, as VE mainly depends on speed of the engine, valve area, valve lift, timing of the opening or closing




446

of valves and residual gas fraction rather than on load variation. Rama Mohan and Murali Krishna also observed the similar trends. VE increased in CE and decreased in LHR engine at their respective optimum injection timings, compared with CE at the recommended injection timing. This is due to decrease of unburnt fuel fraction in the cylinder leading to increase in VE in CE while heating effect of the air causes to decrease the volumetric efficiency in LHR engine. However, the VE is marginally more with LHR engine at its optimized injection timing compared to the same version of the engine at the recommended injection timing. This is due to reduction of gas temperatures. VE increased marginally with the advancing of the injection timing and with the increase of injection pressure in both versions of the engine. This is due to better fuel spray characteristics and evaporation at higher injection pressures leading to marginal increase of VE. This is also due to the reduction of residual fraction of the fuel, with the increase of injection pressure. The reason for improving the VE of LHR engine is reduction of gas temperatures.

3.1 Pollution Levels

The variation in the magnitude of smoke intensity, with BMEP, in CE and LHR engine with pure diesel, at the recommend and optimum injection timings at an injection pressure of 190 bars, is shown in Fig.8. At recommended injection timing and pressure, increase of smoke intensity is observed in LHR engine, when compared with CE. This is due to the decreased oxidation rate of soot in relation to soot formation. Higher surface temperatures of LHR engine aided this process. LHR engine shorten the delay period, which increases thermal cracking, responsible for soot formation. Higher temperature of LHR engine produced increased rates of both soot formation and burn up. The reduction in VE and airfuel ratio are responsible factors for increasing smoke levels in LHR engine near the peak load operation of the engine. As expected, smoke increased in LHR engine because of higher temperatures and improper utilization of the fuel consequent upon predominant diffusion combustion. The magnitude of smoke levels is less for entire load range in both versions of the engine, at their respective optimum injection timings, when compared with CE at the recommended injection timing. This is due to increase of air fuel ratios, causing effective combustion in both versions of the engine at their respective optimum injection timings. Many researchers confirmed this data with CE with the advanced injection timing. Higher combustion temperatures are also conducive for reducing smoke levels with CE at its optimum injection timing. Fuel cracking reactions are eliminated with LHR engine due to low combustion temperatures. This confirmed improvement in fuel utilization with the injection timing of 32 o bTDC, which is further confirmed by increased BTE as seen in Fig.4. Dhinagar et al., Rama Mohan] and Murali Krishna also observed the similar trends. At optimum injection timing, smoke levels are observed to be less in CE, when compared with LHR engine at all loads. This is due to improvement in airfuel ratios and higher VE of CE, when compared with LHR engine. Smoke levels decreased with the advancing of the injection timing and with increase of injection pressure, in both versions of the engine, due to better fuel spray characteristics. The variation in the magnitude of NOx levels, with BMEP, in the CE and LHR engine, with pure diesel, at the recommend and optimum injection timings and at an injection pressure of 190 bars, is shown in Fig.9. For both versions of the engine, NOx concentrations raised steadily as the fuel/air ratio increased with increasing BMEP, at constant injection timing. LHR engine recorded higher NOx at all loads when compared with CE. It is due to the reduction of




447

fuelair equivalence ratio with LHR engine, which is approaching to the stoichiometric ratio, causing more NOx concentrations. Increase of NOx emissions are observed, at all loads in CE, when the injection timing is advanced to its optimum value, when compared with CE at the recommended injection timing. Increasing the injection advance resulted in higher combustion temperatures and increase of resident time leading to produce more NOx concentration in the exhaust of CE at its optimum injection timing. Many researchers confirmed this trend. At the optimum injection timing, LHR engine produced lower NOx emissions, at all loads when compared with the same version of the engine at the recommended injection timing. This is due to decrease of combustion temperatures. However, NOx emissions are marginally lower in LHR engine when compared with CE at their respective optimum injection timings, due to increase of residence time in CE. NOx levels increased in CE while they decreased in LHR engine with the advancing of the injection timing. Many other researchers confirmed this data with CE. However, increase of injection pressure increased NOx emissions in CE while they decreased in LHR engine. This is because of decrease of gas temperatures in the LHR engine and increase of the same in CE with the increase of injection pressure.

Table 1 presents the comparison on the magnitudes of PP, MRPR, TOPP and MRPR with the injection timing and injection pressure, at the peak load operation of CE and LHR engine in comparison with CE at 27 o bTDC. The peak pressures at an injection timing of 27 o bTDC are lower in LHR engine in comparison with CE. This is because the LHR engine exhibited higher temperatures of combustion chamber walls leading to continuation of combustion, giving peak pressures away from TDC. However, this phenomenon is nullified with an injection timing of 32 o bTDC on the same LHR engine because of reduced temperature of combustion chamber walls thus bringing the peak pressures closure to TDC. The magnitude of PP increased with advancing of the injection timings and with the increase of injection pressures, in both versions of the engine. Other researchers also observed the same trend with CE.

Table1: Variation of pp, mrpr,topp and tomrpr with injection timing and injection pressure at the peak load operation of ce and lhr engine with pure diesel operation

At the respective optimum injection timings, the magnitude of PP is more in the CE compared with the LHR engine. MRPR increased with the advancing of the injection timing and with increase of injection pressure with both versions of the engine. Other researchers

PP(bar) MRPR (Bar/deg) TOPP (Deg) TOMRPR (Deg)

Injection pressure (Bar)




Injection timing ( o bTDC)

Engine version

190 270 190 270 190 270 190 270 CE 50.4 53.5 3.1 3.4 9 8 0 0

27 LHR 46.1 51.1 2.7 2.9 11 9 0 0 CE 55.5 62.2 3.3 3.8 8 9 0 0 32 LHR 56.5 58.3 3.5 3.7 9 8 0 0 CE 62.2 61.9 3.8 3.4 8 10 0 1

33 LHR 54.5 56.4 3.4 3.6 10 9 0 0




448

also observed the similar trends with the CE. At the respective optimized injection timings, MRPR is higher in CE, when compared with LHR engine. This is because of increased advancing timing with CE leading to more accumulation of the fuel in the combustion chamber.

From the Table1, it can be noticed that the magnitude of TOPP decreased (shifted towards TDC) with the advancing of the injection timing and with increasing of injection pressure in both versions of the engine. This is confirmed that both versions of the engine showed improvement in performance, when the injection timings are advanced to their respective optimum values. There is no change observed in the magnitude of TOMRPR at different operating conditions of the engine.

3.2. Combustion Modeling

The input data on the surface temperatures of the piston needed in the model are chosen from the temperature predicted from the finite element analysis from the reference Figure 10 shows the comparison of typical pressurecrank angle data predicted from the multizone combustion model and the experimental values obtained from the special Pθ software package and TDC encoder for the case of CE and LHR engine at the recommended injection timing and at the recommended injection pressure. It can be seen from this figure that the LHR engine exhibited lower peak pressures in comparison with CE. The above trend indicated that there is larger amount of energy contained in the combustion gases in case of LHR engine which is also indirectly confirmed by the higher surface temperatures estimated in the finite element analysis which formed the input data for the combustion model. CE and LHR engine at the respective optimum injection timings exhibited higher magnitudes of peak pressures in comparison with CE at the recommend injection timing for both cases of experimental and computer predictions. It could be seen clearly that data predicted from the combustion model gave higher values of peak pressures in CE as well as LHR engine in comparison with the experimental results. This is because of the idealized assumptions assumed in the model, which may not exist in reality.

A deviation of 7% is observed between the experimental results and theoretical values, while Rama Mohan reported a deviation of 6.5% and Murali Krishna 7%. Figure.11 sows the typical diagram of variation of air zone temperature and gas zone temperature with respect to crank angle for both versions of the engine at the recommended and optimum injection timings at the recommended injection pressure. From this figure, it could be clearly seen that higher magnitude of gas temperatures are predicted in LHR engine over that of CE at the recommended injection timing. When the injection timing is advanced to the respective optimum values, computer predictions showed that gas temperatures decreased in the LHR engine while they increased in CE. The decrease of gas temperatures indicated saving of waste heat of the exhaust while converting the same into useful work. This is confirmed from the increased BTE observed with advancing of the injection timing with the LHR engine. This is also further confirmed by the reduction of coolant load as seen and reduction of NOx levels with the advanced injection timing for LHR engine. Coolant load and NOx emissions increased in CE with the advanced injection timings, which indirectly proved the validation of computer predicted gas temperatures in CE. Figure 12 shows the variation of air zone temperature and gas zone temperature with respect to the crank angle for CE and LHR engine




449

at the recommended injection timing and at an injection pressure of 270 bar. Higher gas temperatures are predicted in CE, while lower gas temperatures are predicted in the LHR engine with the increase of injection pressure. Increase of coolant load and increase of NOx emissions are confirming the validation of computer predicted gas temperatures in CE, when the injection pressure is increased. These Tables also indirectly confirmed the computer predicted results of gas temperatures in LHR engine with the observation of coolant loads and NOx emissions in LHR engine with an increase of injection pressure.

4. Conclusions

The optimum injection timing is found to be 33 o bTDC for CE, while it is 32 o bTDC for LHR engine at an injection pressure of 190 bar. BSFC at peak load operation decreased by 12%, peak BTE increased by 7%, EGT increased by 20 o C, coolant load decreased by 12%,VE decreased by 8%, smoke levels decreased by 6% and NOx levels increased by 41% with LHR engine at its optimum injection timing and an injection pressure of 190 bar in comparison with CE with pure diesel operation.

Acknowledgments

Authors thank authorities of Chaitanya Bharathi Institute of Technology, Hyderabad for providing facilities for carrying out research work. Financial assistance provided by All India Council for Technical Education (AICTE), New Delhi, is greatly acknowledged.

5. References

1. Krishnan, D.B., Raman, N., Narayanaswamy, K.K and Rohtagi, P.K., “Performance of an AlSi graphite particle composite piston in a diesel engine”, Transactions of Wear, Volume60, pp 205215, 1980.

2. Wade, W.R., Havstad, H.,Ounsted, E.J.,Trinker, F.H. and Garvin, I.J., “Fuel economy opportunities with an uncooled diesel engine”, SAE Paper No. 841286,1984.

3. Woschni, G., Spindler,W. and Kolesa, K.., “Heat insulation of combustion chamber walls–A measure to decrease the fuel consumption of I.C. Engines”, SAE Paper No. 870339, 1987.

4. Cole, R.M. and Alkidas, A.C., “Evaluation of an AirGapInsulated piston in a divided chamber diesel engine”, SAE Paper No. No. 850359, 1985.

5. Parker, D.A. and Dennison, G.M., “The development of an air gap insulated piston”, SAE Paper No. 870652, 1987.

6. Rama Mohan, K., “Performance evaluation of air gap insulated diesel engine with pure diesel”, PhD Thesis, Kakatiya University, 1995.




450

7. Rama Mohan, K., Vara Prasad, C.M. and Murali Krishna, M.V.S., “Performance of a low heat rejection diesel engine with air gap insulated piston”, ASME Journal of Engineering for Gas Turbines and Power, Volume121, July, 1999, pp 530540.

8. Jabez Dhinagar, S., Nagalingam, B. and Gopala Krishnan, K.V., “A comparative study of the performance of a low heat rejection engine with four different levels of insulation”, Proc. of IV International Conference on Small Engines and Fuels, pp: 121126, Chang Mai, Thailand, 1993.

9. Murali Krishna, M.V.S., “Performance evaluation of low heat rejection diesel engine with alternate fuels”, PhD Thesis, J.N.T. University, Hyderabad, 2004.

10. Siegla Donald, C. and Alkidas Alex, C., “Evaluation of the potential of lowheat rejection diesel engine to meet future EPA heavyduty emission standards”, SAE Paper No. 890291, 1989.

11. Miyairi, Y., Matsunisa, T., Ozama,T., Oikawa,,H. and Narasimma, N., “Selective heat insulation if combustion chamber walls for a DI diesel engine with monolithilc ceramics”, SAE Paper No. 890141,1989.

12. Jabez Dhinagar, S., Nagalingam, B. and Gopala Krishnan, K.V., “ Use of alcohol as the sole fuel in an adiabatic diesel engine”, Proc. of XI National Conference on IC Engines and Combustion, pp 277289, I.I.T., Madras, 1989.

13. Miyairi, Y., “Computer simulation of an LHR direct injection diesel engine” SAE Paper No. 880187, 1988.

14. Rafiqul Islam, MD., Subramanyam, J.P.and Gajendra Babu, M.K., “ A thermodynamic computer simulation model for heat rejection direct injection diesel engine”, Proc. of X National Conference on I.C Engines and Combustion, Rajkot, 1987.

15. Hirayasu, H., “Simulation program to predict direct injection diesel engine efficiency and pollutant emissions”, A research report, Department of Mechanical Engineering, Hiroshima University, 1980.

16. Irvin Glassman, “Combustion”, Academic Press, 1977.

17. Anderson, D., “Rate of heat release in diesel engine”, Report DOTTSCOST7756, University of Southampton, England, 1977.

18. Annand, W.J.D., “Heat transfer in the cylinder of reciprocating internal combustion engines”, Proc. of Institution of mechanical engineers, Volume177, No.36, 1963.

19. Panchapakeshan, N.R., Gopala Krishanan, K.V. and Murthy, B.S., “Factors that improve the performance of an ethanoldiesel oil dualfuel engine”, Proc. of II International Symposium on Alcohol Fuel Technology”, Wolfsburg, W. Germany, 1977.




451

20. Rajan, K. and Poonia, M.P., “Experimental investigations of the effect of injection opening pressure, injection timing and compression ratio on the performance and emission characteristics of a diesel engine under idling conditions ”, Proc. of XV National conference on IC Engines and combustion, pp: 215221,Chennai, 1997.

21. Chandorkar, S.B., Dani, A.D. and Lakshminarayana , P.A., “ Effects of injection parameters, fuel quality and ambient on the ignition delay and location of flame kernel in a diesel spray in a quiesent chamber”, SAE Transactions, Paper No. 881227,1988.

22. Greeves, G. and Keehan ,J.O., “ Measurement of instantaneous soot concentration in a diesel engine combustion chamber”, Paper C88/75, pp:7382, Institution of Mechanical Engineers, London, 1975.

23. Haynes, B.S., and Wagner, H.G., “Soot formation”, Energy Combustion Science, Volume7, pp: 229273,1981.

24. Dent, J.C., Mehta, P.S. and Swan, J., “A predictive model for automotive direct injection diesel engine performance and smoke emissions”, Paper C126/82, International Conference on Diesel Engines for Passenger Cars and Light Duty Vehicles, Institution of Mechanical Engineers, London, England, 1982.

25. Aoyagi, Y., Kamimoto,T., Matusi, Y. and Matsuoka,S., “ A gas sampling study on the formation processes of soot and NO in direct injection diesel engine”, SAE Paper No. 800254, Transactions of SAE, Volume89,1980.

26. Plee, S.L., Ahmad, T., and Myers, J.P.,“ Diesel NOx emissions, A simple correlation techniques for intake air effects”, Proc. of IXX International Symposium on Combustion, pp : 14951502, The Combustion Institute, Pittsburgh, 1983.

27. Vioculescu, I.A., and Borman, G.L., “An experimental studies of diesel engine cylinder averaged NOx histories”, SAE Paper No. 780228, Transactions of SAE, Volume89, 1980.

28. Jagadeesan, T.R. and Mathu, S., “Analytical and experimental investigations of ethanol diesel dual fuel combustion system”, Proc. of .VI International Symposium on Alcohol Fuel Technology, pp :194201,Canada ,May,1984.

29. Submally, Ravi., Subrahmanyam,J.P. and Gajendra Babu, M.K., “A combustion model for a dual fuel direct injection compression ignition engine”, Proc. of IX National Conference on I.C. Engines and Combustion, Indian Institute of Petroleum, Dehradun,1985.

30. Prasad, Y.N., Theoretical and experimental investigations on the performance and emission characteristics of CI engine with biogas substitution”, PhD thesis, I.I.T., Delhi, 1987.

performance evaluation of low heat rejection diesel engine ...performance of the engine with superni...

Documents