Experimental Study of The Influence of Crankcase Thermodynamics on In-Cylinder Delivery Ratio

Invention Journal of Research Technology in Engineering & Management (IJRTEM) ISSN: 2455-3689 www.Ijrtem. com Volume 3 Issue 6 Ç September-October 2019 Ç PP 57-71

Experimental Study of The Influence of Crankcase Thermodynamics on In-Cylinder Delivery Ratio 1,

Lawal Muhammed Nasir, 2,Muhammed Nurudeen 1

Faculty of Engineering, University of Abuja, FCT, Abuja Nigeria National Inland Waterways Authority, Lokoja, Kogi State Nigeria

2

ABSTRACT: The power and efficiency of an internal combustion engine is a function of the quantity of fresh charge that arrives the combustion chamber before the closure of the transfer and exhaust pot. The more the quantity in the cylinder the higher the engine performance. Part of the factors that influence this quantity are the thermodynamic properties of the fresh charge in the crankcase. Charge at higher pressure possesses more momentum to flow quickly into the cylinder. In the same vein, increasing the temperature of the fresh charge results in higher kinetic energy and thus faster movement into the combustion chamber. However, the open-down of the transport and exhaust ports together in two stroke, implies that excessively high pressure and temperatures will push out most of the charges during scavenging process, and thus reducing efficiency. This work investigates the influence of the variation of pressure and temperature in the crankcase on the quantity of charge delivered before the closure of transfer ports. A mass flow rate of 0.029 g/s at an engine speed of 3000 rpm with a delivery ratio of 0.92 as the maximum obtainable with a maximum pressure of 0.15 MPa and a temperature of 330 K. I. INTRODUCTION Improving engine performance in two-stroke engine fundamentally requires increase of the air charge into the crankcase. The crankcase is an intermediary unit that receives, stores and conditions incoming air/fresh mixture; for: (1) scavenging of residual gases from the previous cycle and, (2) supply for the current cycle. The thermodynamics of the crankcase fresh mixture comprises of the heat and mass transfer through the unit, the instantaneous state properties of the mixture (like volume, pressure, temperature, and specific heat ratio), and reconditioning of the charge ahead of delivery to the cylinder. The state of the fresh mixture in the crankcase is highly influenced by the dynamics of the upstream and downstream pressure wave resonances in the inlet and exhaust manifolds respectively. One of the important considerations in providing more air/fresh mixture into the crankcase is the frequency fluctuation matching of the pressure wave in the manifolds to the engine speed. In addition, accounting for the port windows mean time-area and the flow resistance offered by the transport systems is crucial to engine performance. Thus, the delivery ratio as one of the scavenging performance indicators, is largely influenced by these factors [1]. The delivery ratio is a performance measure used in two stroke engines in place of volumetric efficiency fundamentally used in four stroke engines. It is defined as the quantity of fresh mixture that arrived the cylinder just before the closure of the transfer port. This charge is delivered in every cycle within a limited time known as the open-down period of the transfer port. Delivery ratio is concerned with the measurement of quantity that arrived the cylinder and not the quantity trapped in the cylinder. Thus, three periods are discretely important: (1) the open down periods of the intake manifold for charge inflow, (2) the isolation period of the crankcase for compression of the trapped charge in the crankcase and (3) the open-down period of the transfer port for scavenging and supply to the cylinder. Despite the low efficiency and high emission issues of two stroke engines, they remain undisputable versatile prime drivers of most domestic and public appliances, as well as small size automotive systems. Their acceptance against four-stroke engines is credited to their simplicity of technology coupled with high power to weight ratio, lower capital and maintenance costs as well as versatility in working position orientations [2, 3, 4, 5].

The Crankcase Control Volume and Transport Systems: The crankcase is bounded by the reed valve, the under-side of the piston and ring, the transport port pipe, the cylindrical piston wall and the remaining parts of the crankcase internal walls as shown in Figure 1. In any given cycle, these parts have direct contact with the air or fresh mixture at one point or the other. During mass transfer, the air or fresh mixture in the manifold or the cylinder as the case may be, become part of the imaginary boundary. It also form part of the displaced volume as a form of mass transfer with its heat content [6].

|Volume 3|Issue 6|

www.ijrtem.com

| 57 |

Experimental Study of The Influence of Crankcase‌ Engine Cylinder CV Exhaust Port Duct Transfer Port Duct

Intake Port Duct

Crankcase CV

Figure 1 The Crankcase Control Volume Source: View adopted from [6].

The crankcase in two stroke spark ignition engine perform three important functions: firstly, it receives and stores required quantity of air/fuel mixture ahead of each consecutive combustion cycle; secondly, it conditions the stored charge to meet the supply requirement of the combustion chamber in terms of thermodynamic properties of the fresh mixture for effective and efficient combustion as well as the dynamics of flow for effective and efficient scavenging. In this respect, many studies have been done: and facts, processes and methods established.

Nagao and Shimamoto [7] investigated the impact of crankcase capacity and the geometric dimension of the intake manifold on the amount of charge delivered to the cylinder in two-stroke engines with delivery ratio as a performance measure. To match period of fluctuation of pressure wave in the entire control volume to the operational speed and open down periods of both the intake and transfer ports of the crankcase, the crankcase volume and the length of the inlet and outlet ducts were varied experimentally. The individual and combined effects of the two units on delivery ratio the chief determinants of delivery ratio in two-stroke engines is the forceful influence of the constraints at the exhaust and intake units which was found to be a function of the crankcase volume. Another work conducted by Spitsov [8, 9] based on optimal volume of crankcase for effective engine heat transfer, revealed that larger crankcase volume is required for better engine cooling. This is only acceptable to some optimal extent as volume increase reduces pressure required to effects scavenging processes and hence lower delivery ratio. It is also shown by [10] that to gain complete use of the impact of the intake manifold, the transfer port window must be extremely large as much as possible. However extremely large windows has two mutually exclusive negative effects of either reducing the engine compression ration wobbling of the rings into the windows.

Another research indicated that increase in crankcase volume results to drop in delivery ratio which to some reasonable extent can be balanced by tuning the intake manifold. This of course will lead to limitation in the versatility of two stroke engines as continual increase in intake or exhaust manifolds leads bulkiness and preset working orientation. Most of the researchers justified their use of delivery ratio against speed as a performance measure based on the fact that holding the throttle aperture at fixed position will cause approximately constant rate of fuel supply. Thus, in-cylinder delivery is being influenced only by the conditions of operation of the engine.

It was concluded in all cases that any decrease of deliver ratio caused as a result of increase in the volume of the crankcase, can be reasonably overcome by the tuning effects of the intake and exhaust manifolds. The relationship between delivery ratio and engine speed were studied [11], with the objective of maximizing delivery ratio on a scavenged two-stroke engine by varying the mean effective area of the intake port. The result of the investigation reveals that delivery ratio decreases hyperbolically as the intake mark number approaches its limiting value.

Other area of interest is the impact of delivery ratio on the cylinder pressure in the occurrence of blow-by, which was investigated. The researcher reported that the pressure in the combustion chamber at the commencement of blow-down significantly varies the delivery ratio produces [12]. An experimental investigation of exhaust temperature and delivery ratio effects on emissions and performance of a two-stroke engine running on gasoline – ethanol, under different load and speed conditions was reported that addition of 5%, 10%, and 15% ethanol by volume showed a substantial increase in delivery ratio and with 35% [13] and 30% reductions in CO and HC emissions respectively [14]. The temperature of exhaust gas and excess air coefficient were measured in an experiment to investigate the performance of an eight cylinder gasoline engine based on engine speed [15] The researchers reported that as engine speed increases, the amount of fuel entering the combustion chamber rises considerably relative to the quantity of air supplied and thus causing reduction in the excess air coefficient. This in turn, lowers the delivery ratio, considering the air fuel ratio (14.7:1). Thus, high increase in fuel without

|Volume 3|Issue 6|

www.ijrtem.com

| 58 |

Experimental Study of The Influence of Crankcase… corresponding increase in air reduces engine performance and increases emission.Gordon [1] presented a historical background of crankcase design with auxiliary air pumps for scavenging with lots of disadvantages in weight and cost of manufacturing. His revelation shows the concept of using crankcase for scavenging with premixed fuel-lubricant is about half the cost of auxiliary scavenging, and at times more expensive than fourstroke of the same power rating. Gordon, thus, concluded that cost remains one of the nucleus reasons for utilizing the crankcase as means of scavenging, hence the need for continuous redesign for improving crankcase scavenging efficiency. In another report, Gordon [1] further established that the frequency of the resonance in the intake system – intake pipe and crankcase, increases as the square root of the aperture of intake manifold; but reduces as the square root of the length of intake manifold and capacity of the crankcase. It was also reported in [7] that attainment of maximum delivery ratio requires the length of the intake manifold to be in the range that support three-quarter of pressure fluctuations to fall within the open down period of the intake to the crankcase. All these reports thus suggest that since high resonance frequency and amplitude are necessary for increased deliver ratio, larger area of flow is more favoured than larger storage area in the form of crankcase. From thermodynamics point of view, the crankcase assists in preheating the supplied charge ahead of its delivery into the combustion chamber.

This improves the starting condition with less energy required to start of combustion [16]. This temperature condition can be viewed as having similar effect as that of the exhaust gas recirculation system. The increase in heat content is due to resistance offered as the charges enter the crankcase and that due to the crankcase compression as the piston reciprocates. The sonic speed in the crankcase is dependent on temperature whose average value can only be assumed since a surge of gases is simultaneously cooled by expanding highly evaporative gasoline and at the same time takes heat from the container of the control volume (which is conventionally made of aluminium alloys).

In addition to the difficulty of tuning the intake and exhaust manifolds, the crankcase volume continually changes with respect to the piston position. The intake manifold is not in any case a simple pipe; as the flow direction, wall contours and thickness, cross-sectional area and other profiles varies through its path. Beyond that, the momentum of the inflow and outflow gases in the intake manifold could also influences the estimation of crankcase control volume. Thus, it becomes imperative to involve experimental means of accomplishing the same thing, by first removing the impact of the lengths of the intake and exhaust manifolds on the output and efficiency. The result from such experiments differs for different categories of engine based on their cylinder capacity. It was concluded by [1], that for small cylinder capacity an exhaust manifold length of 88.9mm taken between port window and the terminus of the exhaust stub is adequate. Medium cylinder capacity engines should receive about 101.6mm stubexhaust, and 114.3mm for others with as high as between 0.35-0.4lts cylinder capacities. These lengths given by Gordon are functions of the aperture and open down periods of the intake, as well as the capacity of the crankcase.

In conclusion, Gordon’s report demonstrates: one, the possibility of fixing the intake open down period for a range of engine speed by changing the length of the manifold; and two, varying open down period for a given length to produce multiple peaks, or a given peak for multiple effective lengths. The reports of [1, 7] on relationship between a port-area and engine speed show that maximum air delivery occur at a particular speed for a given port crosssectional area. The speed at maximum delivery ratio decreases with increase in crankcase volume. Thus, the crankcase pumping effect is approximately the same peaking at engine speed whose value is inversely dependent of the crankcase capacity. In practice, changes compression ratio of the crankcase will result to corresponding change in the strength of resonance in the intake manifold, and it directly affects the ramming effect of the crankcase charge. Thus, the conflicting demands of capacity and strength of resonance can be controlled with the use of crankcase compression ratios which conventionally ranges from 1.35 to 1.55. For hand held engines, the compression ratio is between 1.2 and 1.4 while for medium engines it stands between 1.4 and 1.5, and for high speed engines it is usually slightly higher than 1.5. For the purpose of this work, considered mainly for medium engines, a mean value of 1.45 is adopted as a starting value for initial boundary condition.

The delivery ratio is maximum when the number of pressure fluctuations is approximately 3, as reported in [1].

This is driven by engine speed and exhaust port area and the exhaust manifold length. The parameters that can easily be controlled during design are the area and length, while the control of speed is based on application and safety require in terms of emission. Flow of fluids occurs due to pressure differences. In compressible flow, the mass flow rate becomes limited or choked when the velocity of flow is sonic. In which case, the flow becomes independent of the downstream condition and purely depends on the upstream thermodynamic properties – density, ρ0 , and specific heat ratio, k.

|Volume 3|Issue 6|

www.ijrtem.com

| 59 |

Experimental Study of The Influence of Crankcase‌ II.

MATERIALS AND METHODS

To establish the influence of thermodynamics properties of the crankcase control volume on the delivery ratio of the combustion chamber of two-stroke engine, the law of conservation, momentum and energy are used. The first approach was the establishment of the instantaneous control volume. The crankcase volume is made up of three parts, namely: displaced volume, clearance volume and transfer tube volume. The schematic diagram in Figure 3 presents the mass and heat transfer across the crankcase control volume.

VC

TDC

VS

x S BDC x

VCC TPO = 118.5o

Figure 2 Mass and heat transfer across the crankcase control volume The displaced volume ideally, is equal to the cylinder swept volume [1]. Thus, the instantaneous crankcase volume Vcc (θ) from Figure 2, is given as: Vckc (θ) = Vclr |BDC + Vd (θ) + Vtt (1) The clearance volume is chosen based on [1], who reported that the optimum crankcase compression ratio rcc , is approximately 1.41. Thus: V | +V | rcc = clr BDC | cc BDC = 1.41 Vclr BDC

(2) V | Where the clearance volume, Vclr |BDC = cc BDC, is the crankcase clearance volume when the piston is at BDC in 0.41 m3, Vcc |BDC is the crankcase volume displaced when the piston at BDC in m3 (= Vd ) and the sum of the two, (Vcc |clr + VCC |BDC ) is the crankcase total volume.

Since the diameter Dc, of the cylinder is constant, the variation in volume is principally a function of the position of the piston đ?‘§(đ?œƒ), along the stroke, and is given as: z(θ) = l + r(1 − cos ( θ)) − √(l2 − r 2 sin2 ( θ) (3) where r is the crankshaft off-set and l is the length of the connecting rod.

Modification of crankcase case control volume :The modification of crankcase volume is carried out by removing the bottom template and fixing several sheets, n, of gaskets of constant thickness t; before covering up the bottom side. This will allow constant increase in volume. The top cylinder is also remove and the stroke modified by putting some gaskets one after the order. The schematic diagram for the modified cylinder stroke and crankcase volume is shown in Figure 4x.

TDC E

S t

BDC R

D

b

t

Figure 3 Modified crankcase and cylinder volumes to optimize delivery ratio

|Volume 3|Issue 6|

www.ijrtem.com

| 60 |

Experimental Study of The Influence of Crankcase‌ Geometric Formulation of the Crankcase Control Volume : As presented in Figure 3.6, the total control volumes of the modified combustion chamber volume is given in Equation (3.6), to be: 0, at TDC đ?œ‹đ??ˇ2 đ?‘‰đ?‘‡ = đ?‘‰đ?‘? + đ?‘‰đ?‘‘ = đ?‘‰đ?‘? ∗ đ?‘§(đ?œƒ) = { (4) 4 đ?‘§đ?‘šđ?‘Žđ?‘Ľ , at BDC The compression ratio as well as the port timing and open-down duration, are dependent on the clearance volume in the two control volumes. With spacer gaskets of thickness t, the volumes of the cylinder and the crankcase will be: Ď€D2

VT = (Vc + Vd ) + *nt; 4 (5) where n, is the number of metal gaskets used t, is the thickness of each metal gaskets applied the corresponding total crankcase volume is: πd2

Vckc,T = (Vclr |BDC + Vd + Vtt ) + ∗ nt 4 where đ?‘‰ckc is the crankcase clearance volume Thus, the engine geometric compression ratio, đ??śđ?‘…đ??ş is:

(6)

đ?œ‹đ??ˇ2

(đ?‘‰đ?‘? + 4 ∗đ?‘›đ?‘Ą)+đ?‘‰đ?‘‘

đ??śđ?‘…đ??ş =

đ?œ‹đ??ˇ2 ∗đ?‘›đ?‘Ą) 4

(đ?‘‰đ?‘? +

;

(7a)

While the crankcase compression ratio, đ??śđ?‘…đ?‘?đ?‘&#x; is: đ?œ‹đ?‘?2

đ??śđ?‘…đ?‘?đ?‘&#x; =

(đ?‘‰đ?‘?đ?‘? + 4 ∗đ?‘›đ?‘Ą+Vtt )+đ?‘‰đ?‘‘

(7b)

đ?œ‹đ?‘?2 ∗đ?‘›đ?‘Ą+Vtt ) 4

(đ?‘‰đ?‘?đ?‘? +

As for port timing, the opening and closing occurring at crank angle, đ?œƒ, correspond to a vertical displacement đ?‘§(∆đ?œƒ) đ?‘›đ?‘Ą đ?‘§(đ?œƒ), where increase in angular displacement ∆đ?œƒđ?‘§ is: ∆đ?œƒđ?‘§ = ∗ 180đ?‘œ = ∗ 180đ?‘œ đ?‘†

And the corresponding increase in open-down duration, ∆đ?œƒđ?‘œđ?‘‘ is ∆θod =

nt S

�

*360o .

The Instantaneous Thermodynamic State of Crankcase Charge : Applying the equation of state to the 1 dVcc 1 dPcc 1 dTcc 1 dR crankcase: [đ?‘ƒđ?‘‰ = đ?‘…đ?‘‡]đ?‘?đ?‘? , is differentiated with respect to time, t, give, + = 0. At Vcc dt

Pcc dt

Tcc dt

R dt

speeds higher than 700 rpm, the flow rate is assumed fast enough to consider the effect of blow-by gas to be negligible in temperature but not in mass. Thus, crankcase it is assumed that flow across the crankcase is adiabatic. �� That is, = 0. thus, the derivative of the ideal gas equation becomes: k dVcc Vcc dt

đ?‘‘đ?‘Ą

+

1 dPcc

Pcc dt

-

k dR R dt

= 0.

(8)

In terms of the instantaneous crank angle, θ, and the angular velocity đ?œ” = 6đ?‘ (where N is in rpm), Equations (12) and (15) is linked together to give influx into the crankcase as: k k dVcc Vcc dθ

+

1 dPcc P dθ

=

2kp Ď

6N

o * √(k-1)Ď

k

2kPcc Ď

o

Cd A V

*√1- (

Pcc (k-1)/k po

)

pcyl 1/k

Cd A

for inflow

pcyl (k-1)/k

-( √ * * ( ) *√( ) pcc pcc { 6N (k-1)Ď cc V From which the instantaneous crankcase pressure is given as: Pcc(i) ∗ (1 + [ Pcc,(i+1) = { The term dVcc dθ

=

Pcc(i) ∗ (1 − [

dVcc

k 6N

k

2kp Ď

√ o i ∗ 6N (k−1)Ď o

2kPcc Ď

√(k−1)Ď âˆ— cc

Cd A V

Cd A V

∗(

∗ √1 − (

pcyl 1/k pcc

)

(9) -1) for outflow

Pcc(i) (k−1)/k po

∗ √(

)

−

pcyl (k−1)/k pcc

)

k dVcc Vcc dθ

−1+

])

inflow (10)

k dVcc Vcc dθ

]) outflow

, on the right-hand side is defined by a backward difference as:

dθ Vcc(i) −Vcc(i−1)

(11)

θ(i) −θ(i−1)

Mass flow Rate as a Function of Thermodynamic Properties of Fresh Charge : The mass flow into the crankcase are made up of fresh mixture of air and fuel from the intake manifold, as well as blow-by influx from the combustion chamber plus left-over from the previous cycle. Thus,

|Volume 3|Issue 6|

www.ijrtem.com

| 61 |

Experimental Study of The Influence of Crankcase‌ mcyl = mmf + mbb + mcc . For a given crankcase volume with known upstream stagnation properties, the flow rate through an intake valve or port into the crankcase and out of the crankcase is described by the equation of compressible flow through a restriction, as: 1/k

P

ṁi = Cd *A*Ď 0 *c0 * ( 0 ) T0

*√

2

k-1 k

P

(k-1)

[1- ( i ) ]

(12)

P0

The quantities Ď 0 , P0 andT0 represent the density, pressure and temperature of fresh mixture in the manifold system, while c0 , is its corresponding speed of sound and Pi , as the downstream (crankcase) pressure, while for flow out of the crankcase through the transfer port into the cylinder, Ď 0 , P0 andT0 represent the density, pressure and temperature of fresh mixture in the crankcase, and c0 , is its corresponding speed of sound; Pi , in this case is the cylinder pressure. A and Cđ?‘‘ the flow area and discharge coefficient respectively for each case. For choked k

P

flow, ( i ) ≼ ( P0

đ?‘˜+1 k−1 ) 2

and the flow rate is k+1

đ?‘˜+1 2(k−1) ( ) .

2

ṁmax = Cđ?‘‘ ∗ A ∗ Ď 0 ∗ c0 ∗ (13) The value of the area of flow, A, varies from the point of opening to the point it completely opened. The mean effective area, AE, is used. This is defined as the average area of flow that will represent the entire variation and produce the same mass flow rate as the summation of the flow rate from penning to the closing of the ports. The mean effective area is given as: đ?œƒ 1 đ??´Ě… = ∗ âˆŤđ?œƒ đ?‘–đ?‘?đ?‘? đ??´đ?‘š đ?‘‘đ?œƒ đ?œƒđ?‘–đ?‘?đ?‘? −đ?œƒđ?‘–đ?‘?đ?‘œ

đ?‘–đ?‘?đ?‘œ

(14) Equation 14 is used for simulation while the variation in experiment is accounted for with the adoption of commonly used values of coefficient of discharge Cd. Thus, đ??´Ě… = đ??śđ?‘‘ ∗ đ??´đ?‘Žđ?‘?đ?‘Ąđ?‘˘đ?‘Žđ?‘™ .

The mass flow rate of air-fuel mixture into the crankcase and out of it with respect to the crank angle, θ, and engine speed N in rpm, is expressed as: (Nagao and Shimamoto, 1972): dm dθ

−6N ∗ Cd A ∗ √

2kpĎ

= {

6N ∗ Cd A ∗ √

2kpĎ k−1

∗(

∗ √(

k−1 p pcc

1/k

)

p pcc

)

(k−1)/k

∗ √1 − (

p pcc

− 1, for flows into the crankcase

)

(k−1)/k

(15) ,

for flows out of the crankcase

Delivery Ratio Dependent on Thermodynamic Properties of the Fresh Cahrge : The delivery ratio of a two stroke spark ignition engine is the ratio of quantity of fresh charge delivered to the combustion chamber from the manifold to the swept volume of the combustion chamber. m đ?‘&#x;d = fa Ď fa ∗Vs

(16) The quantity of fresh mixture delivered is affected by three major factors: the type of fuel used and hence fuel/air ratio as well as heat of vaporization; the geometry and operation of the crankcase systems – which comprises the instantaneous state of the control volume, the crankcase compression ratio, the intake and transfer port design; and the engine operation parameters – engine speed, mixture temperature and pressure. These factors are incorporated in the delivery ratio relation vie: ṁfa đ?‘&#x;d = Ď fr ∗Vs N

(17) At any given time, the combustion chamber has three different gases of component mass: mf for the fuel in atomized form, ma for the oxidation agent (air) and mr for the residual gas from the previous cycle. Thus, the mass of the fresh charge in the cylinder, mfa = mf + ma , and the total content of the combustion chamber is: m m mt = mfa + mr . Taking the fraction of residual gas to be: g r = r, and that of the fresh charge to be:g fa = fa; mt

then, the mass of fuel-air, mfa = mt (1 − g r ).

In terms of the combustion geometrical parameters, Vs = VT − VC =VC (rC − 1) The delivery ratio can thus be expressed in terms of the compression ratio and the fuel-air ratio Îś, as: m (1−gr ). rc đ?‘&#x;d = t ∗ Ď fr ∗(1+Îś)

Vc (rc −1)

mt

(18)

In terms of thermodynamics properties the delivery ratio may further be expressed as:

|Volume 3|Issue 6|

www.ijrtem.com

| 62 |

Experimental Study of The Influence of Crankcase‌ đ?‘&#x;d =

Mt Mfa

∗

Tfa,0 Ti

∗

Pi Pfa,0

1

∗ (1+Îś) ∗ [

rc (rc −1)

−

1 k(rc −1)

Pe

∗[

Pi

+ (k − 1)]]

(19)

Flow seizes when differential pressure is zero. This only happened when the ratio

Pe Pi

= 1, so that the entire

property at the point of no flow are those of thermodynamics and are independent of geometric forms. Thus, the delivery ratio dependence on the thermodynamic properties of the crankcase charge is: T M P 1 đ?‘&#x;d = t ∗ fa,0 ∗ i ∗ (1+Îś).

Mfa

Ti

Pfa,0

(20) Initial and boundary conditions: The equations of volume, pressure and mass flow rate are fundamentally those of initial value problems. Thus, the initial conditions of crankshaft angle, volume and gas pressure are; at đ?‘Ą = 0, đ?œƒ − 0, đ?œƒĚ‡ = 0 đ?‘ƒ = 10140 đ?‘ƒđ?‘Ž, T = 290 K.

III.

EXPERIMENTAL SETUP AND PROCEDURE

Equipment Description : Two kinds of engines were utilized. One is a test bench in Bayero University Kano (BUK) – Thermal Laboratory and the other is a Tiger Generator, model TG1400 obtained newly to investigate the possible influence of changes in volume on pressure and temperature in the crankcase. The main experiment for variation of thermodynamic was conducted with the facility in BUK. The two-stroke SIEs are presented pictorially in Plate 1 (a), (b) and (c). The fuel type used is gasoline with about 4.5% engine oil well mixed in the fuel tank. It is then metered via carburetor based on the quantity of air supply – air fuel ratio is quoted to be 15:1, and a metering rate of 1.5 g/s +1%. For the new engine, the carburetor is vertically positioned with a disc diameter of 35.7 mm. The engine’s intake manifold has a capacity of 85 g/s + 15% and a tuned length of 228 mm. the cross sectional diameter of the intake manifold flushes with that of the carburetor i.e. 35.7 mm. Ashcroft Manifold Pressure Gauge that utilizes a Burdon tube incorporated is used. It has a resolution of 0.02 bar and sensitivity of 5 with an accuracy of 0.2%. A Grill-AF120 control unit utilizing an air fuel ratio of 14.7:1 for spark ignition engine is used to drive the pump in the engine from BUK, while the new engine’s fuel is supplied using gravity as its drive. The fuel pump have an average flow rate of 1.25 g/s through the carburetor while the mixture supply to the crankcase is majorly influenced by the vacuum pumping effect of the piston. The crankcase capacity is approximately 38 cl with a compression ratio of 1.45. The cylinder bore for the new engine is 56.5 mm and a stroke of 58.1 mm. the spark advance angle is 18.5o bTDC, and a maximum speed of 3,000 rpm. The BUK device, has a bore of of 0.07396 m, a stroke of 0.07548 m, a spark advance of 26 o , a combustion duration of 56o maximum speed of 9000 rpm and maximum operating pressure and temperature of 25.9 and 1972 K respectively. SAE30 lubricant is mixed with the fuel at a percentage of 4.5% per litre of gasoline. The port angle are: intake port open IPO = 294.5o, intake port close IPC = 65.6o, transfer port open TPO = 18.5o, transfer port close TPC = 242o, exhaust port open EPO = 102o and exhaust port close EPC = 258o.

Table 2.1: Basic Specifications of the Engine Parameter Description Engine Bore [m] Stroke [m] Displacement [m^3] Crankcase Clearance Volume [m^3] Crankcase Total Volume [m^3] Average Secondary Compression Ratio Actual Secondary Compression Ratio Corresponding Vcr/Vd Mass flow rate at 2,500 rpm density of Air Actual Area at 75o CA Flow Velocity Time for Optimum Supply

|Volume 3|Issue 6|

www.ijrtem.com

Values 0.0565 0.05807 0.000145 0.000364 0.000513 1.41 1.40013 2.529412 0.0135 1.177 0.000872 34.12117 0.003132

| 63 |

Experimental Study of The Influence of Crankcase‌

(a) New Engine TG1400 Crankcase

(b) Crankcase Modified from below

(d) Old Engine Bench TD110 from BUK Plate 1 Views of the two-stroke plant used for the Experimentation Experimental Procedure : The experiment was designed to measure the basic thermal properties – pressure, volume and temperature of the air/fuel mixture flowing across the crankcase. A discharge coefficient of 58.2% is used for the intake port reed valve, and 62 % for the transfer port. The open down period was determined by the various port angles and the engine speed. At a given intake pressure with the throttle angle and engine speed as inputs, the pressure growth and mass flow rate of charges into and out of the crankcase were used as outputs to estimate mass flow rate across the crankcase control volume. A total of 15 tests were conducted. In each case, the engine was run for 10 minutes to attain steady operation while the dynamometer is at no load condition. The throttle plate is set at angle to give approximately, the required engine speed. A stop watch is provided to keep the timing equal for each run. At intervals of 5 minutes, the crankcase pressure, temperature and air/fuel mixture flow rate were read from the display unit while the quantity of fuel used was measured directly by subtracting its previous position in the burette for the current one. An average value is computed from three (3) runs. The throttle plate lever was then adjusted to give an engine speed difference of at least 500 rpm. Both the angle and the corresponding speed were recorded. The crankcase pressure variation over each set of cycles (depending on the speed) over 5 minutes were estimated. The corresponding pressure and temperature in the cylinder were obtained from the panel readings. These processes were repeated for engine speeds 1000 rpm up to 7 000 rpm in steps of 500 rpm corresponding to throttle angle of approximately 30o to 75o with increments of 5o per run. The discharge coefficient was obtained from the ratio of pressure in the crankcase to the pressure in the manifold during the open down period of the intake port. For the transfer port, the discharge coefficient is taken as the ratio of the average pressure of the cylinder after the blow-down to the crankcase pressure at the same time. The manifold pressure was measured with MAP sensor and the corresponding crankcase value computed using ideal gas relationship and measure of changes in volume which is easily measurable. On the other hand, a piezoelectric sensor was used for the in-cylinder pressure. The corresponding in-cylinder temperatures were read from the panel display unit. In each case, the intake pressure was always returned to 0.104 MPa before the commencement of a test.

|Volume 3|Issue 6|

www.ijrtem.com

| 64 |

Experimental Study of The Influence of Crankcase‌ IV. RESULTS AND DISCUSSIONS Variation of the crankcase and control volume over a cycle : Figure 4.1 shows the variation of the crankcase and the combustion chamber volumes. The two control volumes have the same profile but exactly opposite. At top dead centre TDC, the swept volume in the combustion chamber is zero leaving only the clearance volume at the top (0.000015 m3), while the crankcase volume is maximum with a value of 0.00045 m 3. At the bottom dead centre BDC, the maximum cylinder volume is 0.00014 m 3 and the corresponding crankcase volume is 0.00032 m3 given a maximum ratio of approximately 2.286. The two control volumes profiles are sinusoidal with the cylinder representing a sine curve and crankcase representing a cosine curve.

Crankcase Clearance Volume

0

100

200 Crank Angle [deg.]

Cylinder

300

400

Crancase

Main Cylinder Clearance Volume

Instantaneous Control Volume [m^3]

Control Volume variations with Crank Angle 0.0000005 0.0000005 0.0000004 0.0000004 0.0000003 0.0000003 0.0000002 0.0000002 0.0000001 0.0000001 0.0000000

Figure 4 Crankcase and Cylinder Control Volume Variation after modification Crankcase and In-Cylinder Pressure Variations over a Cycle: Figure 4 presents the instantaneous pressure of the crankcase (in red), the actual combustion chamber pressure (in blue) and the theoretical combustion chamber (in green) They all represent the profile of 20o ignition timing as obtained from BUK facility. The cylinder pressure is on the primary axis while the crankcase pressure is displayed on the secondary axis. The analysis starts from BDC, indicated as 0 o.

In-Cylinder and Crankcase Pressure Profile over a Cycle for 20o SOC bTDC

3.50

0.26

3.00 2.50

0.21

2.00 0.16

1.50 1.00

0.11

0.50 0.00

Crankcase Pressure [Mpa}

In-Cylinder Pressure [Mpa]

4.00

0.06 0

50

100

150

200

250

300

350

400

Crank angle [deg.]

Experimental

Simulated

Crankcase

Figure 5 Crankcase and Cylinder Pressure variations Figure 5, the pressure in the crankcase remains approximately constant between 0 o (-180o) and 210o. Between these periods the intake port is opened-down, and the piston is predominantly in the upward stroke, thus, making the crankcase pressure to be slightly below atmospheric. This condition is responsible for the influx of the charge

|Volume 3|Issue 6|

www.ijrtem.com

| 65 |

Experimental Study of The Influence of Crankcase‌ mixture. On the down-ward stroke of the piston, compression takes place in the crankcase noticeable at 210 o (30o aTDC) even though, the intake port is still opened down. The pressure then increases approximately linearly to a maximum value of 0.156MPa at TPO 300 o (120o) after which it reduces non-linearly to a minimum value of 0.075MPa, and sets to commence the next cycle. The trends of two in-cylinder pressure plots agrees in phase and values with the results of Christopher (2009) except for peak pressure location which leads by about 4.5o. Ceviz and Kaymaz (2005) reported an MBT timing of 20o with approximately the same peak pressure at a speed of 3,000 rpm and a lower intake pressure of 0.071MPa as against the value in the work (0.104MPa).

The summary of the pressure at important points are given in Table 1.

Table 1 Pressure values at some critical position of the crank angle.

Critical Position

Angular position

Instantaneous Pressure [MPa] Intake Transfer Exhaust Port Port Port

Crankcase Port

TDC

0o/360o

0.040

0.055

0.060

0.054

IPC

65.5o

0.020

-

-

0.090

EPO

102

o

-

-

0.160

-

TPO

118.5o

-

0.140

-

0.156*

o

BDC TPC EPC IPO Minimum

180 242o 258o 294.5o Varies

0.060 0.045 0.015

0.080 0.020 0.018

0.080 0.060 0.040

0.120 0.020 0.020** 0.020

Maximum

Varies

0.080

0.140

0.160

0.160

*Maximum at 118.5o CA

**minimum at 294.25oCA.

From Table 1, the crankcase pressure is minimum (0.02MPa) and quiet below atmospheric at IPO - (294.25o CA), thus, creating vacuum to establish pumping effect. It attains maximum value of 0.16MPa, shortly before EPO at 118.5o. The ratio of transfer pressure to exhaust at opening and closing of transfer port are 1.25 and 0.92 for optimum performance.

Average Charge Temperature [K]

Crankcase Average Temperature Profile: Figure 6 presents the crankcase average temperature profile over a cycle. In (a), the trend begins at BDC marked 0 o, moves up to TDC, 180o and back downwards to BDC again. In (b) the profile is temperature in the combustion chamber is presented for comparison.

Avergae Crankcase Temperature Cycle

over a

340 330 320 310 300 290 280 270 0

100

200

300

400

Crank Angle [deg.] (a) Crankcase temperature trend

|Volume 3|Issue 6|

www.ijrtem.com

| 66 |

Experimental Study of The Influence of Crankcase…

Average Charge Temperature [K]

Avergae In-Cylinder Temperature Cycle

over a

2500 2000 1500 1000 500 0 0

100

200

300

Crank Angle [deg.]

Cylinder

(b) In-cylinder temperature trend Figure 6 Average crankcase and in-cylinder temperature over a cycle From the graphs, at the bottom dead centre BDC (0o) when both the transfer and exhaust ports are opened down, the average temperature in the crankcase is approximate 309 K. At about 32o aBDC the temperature appreciates from 309 K at 25o to 315 at 60o CA. This increase is due to slight back flow from the cylinder to the crankcase.

The temperatures of the two chambers remain the same and constant even when the transfer and exhaust ports were closed, until the intake port was opened at 115o aBDC. The temperature then fell rapidly to 286 K at 120o due to the incoming fresh charge from the manifold. As the piston continues to go up, effective compression starts and temperature begins to rise in the cylinder. Some work is done in moving up the piston – and friction generates some heat, part of which causes the initial rise in crankcase temperature from about 286 K to 330 K. It then remains at this level until the intake port is opened and gases at atmospheric (or manifold) condition came in and force the temperature down to approximately 320K. From this value, it reduces non-linearly to about 314K due to expansion – while the cylinder rises to TDC. On its way back, compression takes place which increases the temperature linearly to supply condition before the transfer port is opened at 118.5o CA. Towards the BDC, there tends to be some flow back from the cylinder when the crankcase pressure becomes lower, thus causing sharp increase in temperature again. However, the downward pushing effect of the under piston keeps the flow from crankcase to cylinder. As against the increase in temperature for pre-mix supply, the direct air supply of similar set up conducted by [17], utilized initial temperature of 300K. This difference of 20K provides a better condition for start of combustion in this study than Philip’s system.

From Table 1, EPO at 118.5oCA corresponds to a crankcase maximum pressure of 156kPa and cylinder volume of VT of 0.00013260m3 at - made of the clearance volume and volume due to piston displacement x. This displaced volume in the cylinder is equal to the volume displaced in the crankcase. The engine tested has an optimized crankcase clearance volume of 0.000368m3, thus, the crankcase displaced volume is 0.000114862m3 for a pressure rise from 0.054 at TDC to 0.156MPa at TPO. Utilizing the optimal crankcase compression ratio - 1.41, the total volume of the crankcase is 0.000514m3. Thus, the available crankcase volume at TPO is, Vcr = 0.000514 0.000115m3. This gives V2 to be 0.000399m3. With an average specific heat ratio k, of 1.3527, as earlier on assumed, the supply temperature of charge to the cylinder at TPO is 332.5 K. For cycle analyses the average crankcase temperature is taken to be the median value (324.5K).

Average Crankcase and In-Cylinder Energy: Figure 7, shows the average crankcase and in-cylinder instantaneous rate of heat transfer over a complete cycle, obtained from five runs of two-stroke Spark Ignition – BUK Facility. The profile is very much like that of the temperature. It gives the complete trend at BDC, TDC and around the mass transfer periods.

|Volume 3|Issue 6|

www.ijrtem.com

| 67 |

Experimental Study of The Influence of Crankcase‌ Mass Flow Rate in and out of the Crankcase at 2,000 rpm

Mass Flow Rate (g/s)

0.02

0.01

0.00

-0.01 95

120

145

170

195

220

245

Crank Angle (deg.)

Out flow

Inflow

Figure 7 Mass flow balances across the crankcase and cylinder control volumes From Figure 7, the highest mass flux does not occur immediately at TPO or EPO, this is because flow rate depends on area, and the area available at the beginning of port opening is very smaller compared to when fully opened – the point at which maximum mass gets into and out of the cylinder is about 72.6 % of the TP window and 67.3 % of the EP window. After this peak position the mass flux decreases slowly and approximately linearly to zero at 241.5o. The result also shows that crankcase mass is always higher than that in the cylinder. During the upward stroke, the IP is opened at 65.5o leading to increase in influx to the crankcase from 9.05*10-5kg to 11.5*10-5 kg.

This mass is retained until TP opens for transfer into the cylinder. On the other hand, as the piston moves up from BDC, the EP is opened down and mass flows out of the cylinder, hence the reduction in the resident mass before the closure of the EP after which the mass will be trapped and retained till about 102o, when EP is opened again for blow-down and massive out-flux.

Flow Rate through the Cylinder at Speed Range of 500 to 5500 rpm : Figure 8 shows the impact of speed on the mass flow rate through the cylinder for a speed range of 500 to 3,000 rpm. The experimental values show approximately a linear relationship with engine speed, but simulated result presented a slightly non-linear trend.

In the two cases, flow rate increases with rpm. Its maximum value is 0.03g/s occurring at 5,000 rpm, and falls with further increase in rpm. This can be explained to be due to delivery overlap for the engine supply and time lag at higher speeds.

Mass Flow Rate [g/s]

Crakcase Out Flux Rate Vs Engine Speed 0.035 0.030 0.025 0.020 0.015 0.010 0.005 0.000 0

1000

Experimental

2000

3000

Engine Speed [rpm] Simulation

4000

5000

6000

Poly. (Experimental)

Figure 8 Flow Rate through the Crankcase to the Cylinder at Speeds of 500 to 5500 rpm

|Volume 3|Issue 6|

www.ijrtem.com

| 68 |

Experimental Study of The Influence of Crankcase‌ On average, mass flow rate increases with increase in rpm. Simulated result (in green) shows that flow rate bears a non-linear relationship with speed, increasing gradually to a peak rate of 0.0325g/s at 5000 rpm, as against experimental result that is approximately linear up to a speed of 3000 rpm and lost its linearity after that. The values are very close with only 3.2% average difference.

Results of Mass Fraction of Fresh against Scavenging Ratio: Figure 9 shows the mass fraction of cylinder contents at different values of delivery ratios. It is important to know what quantity of the content in the combustion chamber will participate in combustion. Part of the mixture has gas from the previous cycle in two ways: (1) from the blow-by into the crankcase, and (2) from the left over after exhaust is closed. The sum total of theses spent gases do not participate and thus, reduces the potency of the fresh charge. From figure 4.6, the fraction of potential fresh charge increases as the delivery ratio increases. Its reaches its maximum value at about a delivery ratio of 0.93, which corresponds to a charge of 82% of the content of the cylinder after the closure of the exhaust port.

In-Cylinder Mass Fraction of Fresh Charge Vs Delivery Ratio at an Engine speed of 3,000 rpm Mass Fraction of Fresh Charge

1.000 0.900 0.800 0.700 0.600 0.500 0.400 0.300 0.200 0.100 0.000 0.30

0.40

0.50

0.60

0.70

0.80

0.90

1.00

Delivery Ratio Ratio Exp. Before

Exp. After

Simulated

Figure 9 Mass fraction of fresh charge against Delivery ratios Figure 10 shows variation of the delivery ratio with temperature ratio. The temperature of the working fluid is normalized by taking ratio a ratio of it to a reference temperature of 273 K.

Variation of Delivery ratio against characteristic Temperature ratio 0.760

Delivery Ratio

0.750 0.740 0.730 0.720 0.710 0.700 0.95

1.00

1.05

1.10

1.15

1.20

Ti/To

Figure 10 Variation of delivery ratio with temperature relative to 273 K

|Volume 3|Issue 6|

www.ijrtem.com

| 69 |

Experimental Study of The Influence of Crankcase…

From the graph of Figure 10, the delivery ratio increases with increase in temperature at a rate of 0.43. the temperature To is chosen to be 273 K. This is reasonable because, with increase in temperature the kinetic energy of the charge increases and thus the pressure and hence the mass flow rate into the cylinder.

V. CONCLUSION It is established in the work that the delivery ratio of a two-stroke engine is influenced by the thermodynamic state of the charge in the crankcase. Variation in pressure, temperature and energy content of the charge causes changes in momentum, kinetic energy and mass and heat transfer. Lower values of these properties results to inadequate propulsion offered for supplying the required charge. On the other hand, extremely high values of these properties cause over excitation of the outflow from the crankcase and much of the charge going into the combustion chamber will get out through the exhaust port un-combusted during the open down periods. As such there is need to investigate the influence of these properties on the quantity of fresh charge that arrives the cylinder at varying conditions of flow. The two control volumes are actually collapsing into one another above and below the piston.

The study is unique as it used a practical approach of altering the volumes of the crankcase and the combustion chamber using some metal sheet gaskets of known thickness and numbers to vary the properties of interest which would have otherwise difficult to achieve. The study is significant because the quantity of power output is dependent on the quantity of charge that arrives the combustion chamber before the closure of the transfer port.

The more the charge the more the power and hence speed and torque, thus improving the efficiency of the engine.

In addition, the effect of the blow-by gas on the thermodynamic properties of the crankcase content is considered.

Thus, it is recommended to install a control system that regulates the thermodynamic properties crankcase charge for optimal delivery into the cylinder.

REFERENCES [1] [2].

[3].

[4] [5] [6]

[7]

[8]

[9] [10]

[11]

[12]

[13]

J. Gordon. Two-stroke tuner’s handbook(2nd ed.).USA,Harry Potter Books, pp.75-90, 2007. [Online] Available: http://edj.net/2stroke/jennings/ Retrieved on October 16, 2018.

M. A. Amin, A. A. Azhar, F. M. S. Mohd and A. L. Zulkarnain. “An experimental study on the influence of EGR rate and fuel octane number on the combustion characteristics of a CAI two-stroke cycle engine”, Applied Thermal Engineering, 71: 248-258, 2014.

V. Pradeep, S. Bakshi and A. Ramesh. “Scavenging port based injection strategies for an LPG fuelled two-stroke spark-ignition engine”, Applied Thermal Engineering, 67: 80-88, 2014.

J. Volckens, D. A. Olson, M. D. Hays, “Carbonaceous species emitted from hand held two-stroke engines”, Energy Procedia, 45: 739 – 748, 2014.

G. Cantore, E. Mattarelli and C. A. Rinaldini. “ A new design concept for 2-Stroke aircraft Diesel engines”, Energy Procedia, 45: 739–748, 2014.

K.T. Jayanth, M. Arun, G. Murugan, R. Mano and J. Venkatesan. Modification of two stroke ignition compression engine to reduce emission and fuel consumption.International Association of Computer Science and Information Technology Journal of Engineering and Technology, 2, pp.23- 42, 2010.

F. Nagao and Y. Shimamoto. The Effect of Crankcase Volume and the Inlet System on the Delivery Ratio of Two-Stroke Cycle Engines. Proceedings of Society of Mechanical Engineers, Internal Combustion Engine Conference, Kyoto, Japan, March 6-9, 1972, 175-193.

J. A. Naser, and A. D. Gosman. Flow prediction in an axisymmetric inlet valve/port assembly using variants of k- ε. Proceeding of Institution of Mechanical EngineersConference on Automobile Engineering,Bellevue, Washington, USA, June 25-28, 1995, 209(4)289-295, DOI:10.1243/PRIME_PROC_1995_209_216_02.

T. J. Dembroski. Piston heat transfer in an air-engine cooled. ERS Metadata. [Online] Available: http://digital.library.wisc.edu/1793/35285 A. Deshmukh, O. Ankaikar, A. Kharote and M. Chinchkar. Effect of Scavenging and Ports on Acceleration of Two Stroke Engine. International Journal of Engineering Trends and Technology (IJETT) – Volume 58 Issue 1- April 2018. http://www.ijettjournal.org M. Ghazikhania D. D. Ganjic. Experimental investigation of exhaust temperature and delivery ratio effect on emissions and performance of a gasoline–ethanol two-stroke engine https://doi.org/10.1016/j.csite.2014.01.001, 2014.

T. Asanuma, and N. Sawa. The effect of length of an intake pipe on the delivery ratio in a small two-stroke cycle engine. Journals of Society of Mechanical Engineers, 3(9), 137-142. 1960.

M. Ghazikhani, M. Hatami, B. Safari and D. D. Ganji. Propultion and power research journal vol 2 issue 4, pp. 276 – 283; 2013.

|Volume 3|Issue 6|

www.ijrtem.com

| 70 |

Experimental Study of The Influence of Crankcase…

[14] [15]

[16]

[17]

E. Ashraf, Engineering science and technology, an international journal, vol. 18, issue 4, pp. 713 – 719.

2015 A. Kilicarslan and Q . Muhammed, Exhaust gas analysis of an eight cylinder gasoline engine based on engine speed.

Energy procedia.

Vol. 110, march, 2017, 459 – 464. https://doi.org/10.1016/j.egypro.2017.03.169. 2017 K.T. Jayanth, M. Arun, G. Murugan, R. Mano and J. Venkatesan. Modification of two stroke ignition compression engine to reduce emission and fuel consumption. International Association of Computer Science and Information Technology Journal of Engineering and Technology, 2, 23- 42, 2010.

J. O. Philip, A numerical investigation of the scavenging flow in a two-stroke engine with passive intake valves. MSci. Thesis submitted to the Department of Mechanical and Materials Engineering, Queen’s University Kingston, Ontario, Canada, 7-29. 2008.

|Volume 3|Issue 6|

www.ijrtem.com

| 71 |