Internal Combustion Engine Technology

**Chapter 1**

**Abstract**

Rightsizing

*Zbigniew J. Sroka*

combustion engine!

fossil fuels [2].

**3**

Work Cycle of Internal

Combustion Engine Due to

It is worth still working on the development of the internal combustion engine, because its time was not yet over. This was demonstrated by the author's review of the literature, indicating at least the perspective of 2050 the universality of the engine as the primary propulsion or support in hybrid transport units. The presented considerations may have a broader perspective, when the thermodynamic problems of a thermal machine such as an internal combustion engine are indicated. This chapter deals with the issues of changing the swept volume known as downsizing/rightsizing. An equivalent swept volume was introduced, defined by the coefficients determining changes in the cylinder diameter and the stroke of the piston. An attempt was made to find the mutual relations to the efficiency of the work cycle and engine operating parameters. The research methodology was proposed as a mix of laboratory tests and theoretical analyses, on the basis of which it was established that while maintaining the same value of the downsizing index, despite the various permissible combinations

of cylinder diameter and piston stroke changes, the cycle efficiency remains

**Keywords:** internal combustion engine, work cycle, rightsizing

**1. Introduction - the essence of the research problem**

support systems for rightsizing geometric changes.

development of internal combustion engines:

unchanged. The engine operating parameters are changing, resulting from the use of

It is the beginning of 2021 and internal combustion engines are not yet dead, although many people predicted their significant reduction in connection with the introduction of hybrid drive in vehicles [1]. And yet this drive still has an internal

When in 2007 the 2nd PTNSS Engine Congress was held in Krakow, in Poland, an international group of scientists and researchers identified three scenarios for the

I.short-term (until 2017): improving the design of internal combustion engines

III.long-term (over 30 years, i.e. over 2037): independence of transport from

to meet ecological standards and the use of alternative fuels,

II.mid-term (2017–2037): development of hybrid systems,

#### **Chapter 1**

## Work Cycle of Internal Combustion Engine Due to Rightsizing

*Zbigniew J. Sroka*

#### **Abstract**

It is worth still working on the development of the internal combustion engine, because its time was not yet over. This was demonstrated by the author's review of the literature, indicating at least the perspective of 2050 the universality of the engine as the primary propulsion or support in hybrid transport units. The presented considerations may have a broader perspective, when the thermodynamic problems of a thermal machine such as an internal combustion engine are indicated. This chapter deals with the issues of changing the swept volume known as downsizing/rightsizing. An equivalent swept volume was introduced, defined by the coefficients determining changes in the cylinder diameter and the stroke of the piston. An attempt was made to find the mutual relations to the efficiency of the work cycle and engine operating parameters. The research methodology was proposed as a mix of laboratory tests and theoretical analyses, on the basis of which it was established that while maintaining the same value of the downsizing index, despite the various permissible combinations of cylinder diameter and piston stroke changes, the cycle efficiency remains unchanged. The engine operating parameters are changing, resulting from the use of support systems for rightsizing geometric changes.

**Keywords:** internal combustion engine, work cycle, rightsizing

#### **1. Introduction - the essence of the research problem**

It is the beginning of 2021 and internal combustion engines are not yet dead, although many people predicted their significant reduction in connection with the introduction of hybrid drive in vehicles [1]. And yet this drive still has an internal combustion engine!

When in 2007 the 2nd PTNSS Engine Congress was held in Krakow, in Poland, an international group of scientists and researchers identified three scenarios for the development of internal combustion engines:

I.short-term (until 2017): improving the design of internal combustion engines to meet ecological standards and the use of alternative fuels,

II.mid-term (2017–2037): development of hybrid systems,

III.long-term (over 30 years, i.e. over 2037): independence of transport from fossil fuels [2].

With the passage of years and the verification of predictions on the basis of real data, the need for the development of internal combustion engines was indicated indirectly in connection with the change from linear to exponential transport index passenger-kilometer, which forces the increase in the production of motor vehicles (passenger cars, trucks and busses) from the current 70 million annually to over 107 million units in 2050 [3, 4].

Finally, in the general discussion, the development scenarios of combustion engines were indicated [16, 17]. In the short-term perspective, i.e. until 2030, the importance of environmental protection was emphasized, and in the longer term, i.e. until 2050, attention was additionally paid to the importance of sustainability

The above considerations have a common denominator - the world does not give

One of the development trends is the downsizing of combustion engines that has

The essence of the research problem presented in this chapter is to demonstrate the existence of parameters describing the engine displacement volume, which is the dominant feature of downsizing/rightsizing, enabling the assessment of the

This means that the main research question can be formulated as follows - is it possible to replace the displacement volume of an internal combustion engine in the considerations on its work cycle, a certain equivalent volume, and can the application of the new solution be used to investigate the cause-and-effect

relationships between thermodynamic parameters and internal combustion engine

tive work cycle for similar values of the downsizing index was assessed.

maintaining or increasing the engine power per liter of displacement.

*peVss*

Indicators with the index "d" indicate downsizing data.

*n*

**2. Rightsizing the internal combustion engine**

downsizing engine power), Eq. (2) is obtained

The search for an answer to the research question is associated with the analysis of the thermodynamic work cycle of the downsized engine. The impact of changes in the swept volume and the equivalent volume on the parameters of the compara-

Development works related to the rightsizing concept are focused primarily on increasing the specific volumetric power. These are therefore actions similar to those previously undertaken for downsizing, when reducing the displacement while

The essence of downsizing results from the power equation, which takes the

*n*

*nd* 30*τ<sup>d</sup>*

<sup>30</sup>*<sup>τ</sup>* (1)

(2)

*Ne* ¼ *peVss*

By changing the engine's displacement according to the rule - the volume "after" is smaller than "before", i.e. *Vssd* < *Vss*, (where: *Vssd* is the engine swept volume after downsizing) and at the same time keeping the engine power *Ned = Ne* (where: *Ned* -

<sup>30</sup>*<sup>τ</sup>* <sup>¼</sup> *pedVssd*

Assuming constancy engine speed *nd* = *n* and constancy of the number of strokes

rightsizing. This trend is not so much about reducing the displacement as it is about choosing the right size in order to achieve a balance between customer expectations for operating comfort and the manufacturer's ability to reduce fuel consumption

up on internal combustion engines. Research centres and universities are still

been going on for over ten years and has recently been modified towards

and safe use of engines in the environment.

*DOI: http://dx.doi.org/10.5772/intechopen.97144*

and CO2 emissions.

performance indicators?

form (1) [18, 19].

*τ<sup>d</sup> = τ*, to give (3)

**5**

working on developing the design of this heat machine.

*Work Cycle of Internal Combustion Engine Due to Rightsizing*

effectiveness of changes in the IC engine operation indicators.

On August 17, 2017, Norman Mayersohn, in The New York Times magazine, in an article entitled "The Internal Combustion Engine Is Not Dead Yet", interviewed Professor John Heywood, the undisputed guru in the design and testing of internal combustion engines. Professor Heywood pointed to the presence of internal combustion engines with a significant share in 2050 – quotation: *"Definitely. John Heywood, a professor of mechanical engineering at the Massachusetts Institute of Technology, predicts that in 2050, 60 percent of light-duty vehicles will still have combustion engines, often working with electric motors in hybrid systems and largely equipped with a turbocharger. Vehicles powered purely by batteries, he estimates, will make up 15 percent of sales"* [5].

In April 2020, the virtual 41st International Vienna Motor Symposium took place (due to the COVID19 coronavirus pandemic), during which the development of internal combustion engines was discussed [6].

It was the time of the "New and Optimized Engines" session, during which Ford presented the latest solutions in the field of EcoBoost technology, emphasizing the importance of charging [7].

Toyota discussed the 1.5 Liter engine solutions from the Toyota New Global Architecture (TNGA) platform, emphasizing the importance of balance between design and application. Among other things, it was discussed: hydraulically variable valve timing, very high compression moderated by Atkinson cycle, longer bore and stroke ratio, application of multi-hole injector system to achieve "high-speed combustion", resulting more than 40% in thermal efficiency [8].

The authors of another presentation mentioned a similar meaning of the modular construction and technological platform for internal combustion engines [9].

The modularity of engines, but in relation to Diesel, was discussed during the session "New SI and CI Engines" [10], where the modular solutions of the BMW company were demonstrated.

Similar to Toyota TNGA solutions at Mercedes-Benz is FAME (Family of Modular Engines), which involves the creation of subsequent engine versions based on the M-254 engine. [11]. Everything is dedicated to the fulfillment of global CO2 fleet targets. Quotation "..*the M 254 paves the way with regard to CO2- neutrality and air quality approaching the sustainability strategy Ambition 2039.*

The importance of the filling process both on the supercharging side and the change of the geometry of the suction system were emphasized. Attention was also paid to the reduction of friction in the piston-cylinder liner system. The summary of the whole was as follows - a quote *… the internal combustion engine is still far from being at the end of the road!*

Environmental protection is the dominant topic in all publications. This is also the case in another study [12], where VW indicated numerous possibilities of meeting Euro 6d standards.

Subsequent studies indicate the importance of alternative fuels, with particular emphasis on hydrogen [13, 14]. The full usefulness of typical hydrogen-powered combustion engines has been demonstrated in relation to the still developed Fuel-Cell technology.

The extensive discussion is not forgotten engine applications truck [15]. Here importance is the durability of use. Considerations were conducted in the perspective of 2050!

#### *Work Cycle of Internal Combustion Engine Due to Rightsizing DOI: http://dx.doi.org/10.5772/intechopen.97144*

With the passage of years and the verification of predictions on the basis of real data, the need for the development of internal combustion engines was indicated indirectly in connection with the change from linear to exponential transport index passenger-kilometer, which forces the increase in the production of motor vehicles (passenger cars, trucks and busses) from the current 70 million annually to over 107

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

On August 17, 2017, Norman Mayersohn, in The New York Times magazine, in an article entitled "The Internal Combustion Engine Is Not Dead Yet", interviewed Professor John Heywood, the undisputed guru in the design and testing of internal combustion engines. Professor Heywood pointed to the presence of internal combustion engines with a significant share in 2050 – quotation: *"Definitely. John Heywood, a professor of mechanical engineering at the Massachusetts Institute of Technology, predicts that in 2050, 60 percent of light-duty vehicles will still have combustion engines, often working with electric motors in hybrid systems and largely equipped with a turbocharger. Vehicles powered purely by batteries, he estimates, will make up 15 percent of*

In April 2020, the virtual 41st International Vienna Motor Symposium took place (due to the COVID19 coronavirus pandemic), during which the development

Toyota discussed the 1.5 Liter engine solutions from the Toyota New Global Architecture (TNGA) platform, emphasizing the importance of balance between design and application. Among other things, it was discussed: hydraulically variable valve timing, very high compression moderated by Atkinson cycle, longer bore and stroke ratio, application of multi-hole injector system to achieve "high-speed com-

The authors of another presentation mentioned a similar meaning of the modular construction and technological platform for internal combustion engines [9]. The modularity of engines, but in relation to Diesel, was discussed during the session "New SI and CI Engines" [10], where the modular solutions of the BMW

Similar to Toyota TNGA solutions at Mercedes-Benz is FAME (Family of Modular Engines), which involves the creation of subsequent engine versions based on the M-254 engine. [11]. Everything is dedicated to the fulfillment of global CO2 fleet targets. Quotation "..*the M 254 paves the way with regard to CO2- neutrality and air*

The importance of the filling process both on the supercharging side and the change of the geometry of the suction system were emphasized. Attention was also paid to the reduction of friction in the piston-cylinder liner system. The summary of the whole was as follows - a quote *… the internal combustion engine is still far from*

Environmental protection is the dominant topic in all publications. This is also

Subsequent studies indicate the importance of alternative fuels, with particular emphasis on hydrogen [13, 14]. The full usefulness of typical hydrogen-powered combustion engines has been demonstrated in relation to the still developed

The extensive discussion is not forgotten engine applications truck [15]. Here

importance is the durability of use. Considerations were conducted in the

the case in another study [12], where VW indicated numerous possibilities of

It was the time of the "New and Optimized Engines" session, during which Ford presented the latest solutions in the field of EcoBoost technology, emphasizing the

of internal combustion engines was discussed [6].

bustion", resulting more than 40% in thermal efficiency [8].

*quality approaching the sustainability strategy Ambition 2039.*

million units in 2050 [3, 4].

importance of charging [7].

company were demonstrated.

*being at the end of the road!*

meeting Euro 6d standards.

Fuel-Cell technology.

perspective of 2050!

**4**

*sales"* [5].

Finally, in the general discussion, the development scenarios of combustion engines were indicated [16, 17]. In the short-term perspective, i.e. until 2030, the importance of environmental protection was emphasized, and in the longer term, i.e. until 2050, attention was additionally paid to the importance of sustainability and safe use of engines in the environment.

The above considerations have a common denominator - the world does not give up on internal combustion engines. Research centres and universities are still working on developing the design of this heat machine.

One of the development trends is the downsizing of combustion engines that has been going on for over ten years and has recently been modified towards rightsizing. This trend is not so much about reducing the displacement as it is about choosing the right size in order to achieve a balance between customer expectations for operating comfort and the manufacturer's ability to reduce fuel consumption and CO2 emissions.

The essence of the research problem presented in this chapter is to demonstrate the existence of parameters describing the engine displacement volume, which is the dominant feature of downsizing/rightsizing, enabling the assessment of the effectiveness of changes in the IC engine operation indicators.

This means that the main research question can be formulated as follows - is it possible to replace the displacement volume of an internal combustion engine in the considerations on its work cycle, a certain equivalent volume, and can the application of the new solution be used to investigate the cause-and-effect relationships between thermodynamic parameters and internal combustion engine performance indicators?

The search for an answer to the research question is associated with the analysis of the thermodynamic work cycle of the downsized engine. The impact of changes in the swept volume and the equivalent volume on the parameters of the comparative work cycle for similar values of the downsizing index was assessed.

#### **2. Rightsizing the internal combustion engine**

Development works related to the rightsizing concept are focused primarily on increasing the specific volumetric power. These are therefore actions similar to those previously undertaken for downsizing, when reducing the displacement while maintaining or increasing the engine power per liter of displacement.

The essence of downsizing results from the power equation, which takes the form (1) [18, 19].

$$N\_{\epsilon} = p\_{\epsilon} V\_{sr} \frac{n}{30\pi} \tag{1}$$

By changing the engine's displacement according to the rule - the volume "after" is smaller than "before", i.e. *Vssd* < *Vss*, (where: *Vssd* is the engine swept volume after downsizing) and at the same time keeping the engine power *Ned = Ne* (where: *Ned* downsizing engine power), Eq. (2) is obtained

$$p\_e V\_{st} \frac{n}{\mathbf{30} \tau} = p\_{ed} V\_{sd} \frac{n\_d}{\mathbf{30} \tau\_d} \tag{2}$$

Indicators with the index "d" indicate downsizing data.

Assuming constancy engine speed *nd* = *n* and constancy of the number of strokes *τ<sup>d</sup> = τ*, to give (3)

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

$$p\_{ed} = p\_e \left(\frac{V\_{gs}}{V\_{sd}}\right) \tag{3}$$

In turn, fuel consumption expressed as the specific value (*ge*) can be written as (4)

$$\mathbf{g}\_{\epsilon} = \frac{1}{\eta\_{\epsilon} W\_{u}} \tag{4}$$

where the useful efficiency *η<sup>e</sup>* is expressed by the relation (5)

$$\eta\_{\epsilon} = \frac{\text{MRL}\_{p} p\_{\epsilon} T\_{o}}{\eta\_{v} W\_{u} p\_{o}} \tag{5}$$

With a reasonable assumption of unchanged value beyond the engine operating after downsizing, useful efficiency becomes dependent only on the brake mean effective pressure (*pe = BMEP*).

Brake specific fuel consumption (*BSFC = ge*) can also be expressed by defining the actual amount of fuel burned per unit time, giving the unit of power (6).

$$\mathbf{g}\_{\epsilon} = \frac{\mathbf{G}\_{\epsilon}}{N\_{\epsilon}} \tag{6}$$

Keeping the constant useful power after downsizing, i.e. *Ned = Ne*, the following Eq. (7) is obtained

$$\frac{G\_{el}}{\mathcal{g}\_{el}} = \frac{G\_{\epsilon}}{\mathcal{g}\_{\epsilon}}\tag{7}$$

of the definition, this indicator shows the change or degree of residue after the

*Wd* <sup>¼</sup> <sup>1</sup> � *AB*<sup>2</sup> dla *<sup>A</sup>* <sup>¼</sup> *Sd*

changes in the swept volume can be distinguished - as in **Figure 2**.

In the graphic interpretation, theoretically and practically three forms of

*Wd* indicator: downsizing and upsizing, important in considering rightsizing. Having knowledge of the design of the combustion chamber and the crank system in the commonly accepted geometric relationships between the cylinder diameter and the stroke [18, 22], as well as based on the actual relationships of these parameters determined on the basis of the engines from the Engine of the Year competition over the years 1999–2019 [19, 20] it was possible to determine the real ranges of variability of the ratio of the cylinder diameter to the piston stroke, which is from 0.77 to 1.30, which results in the value of the *Wd* index in the range of minus

�1.20 on the upsizing side and plus +0.51 in the case of downsizing.

By implementing the idea of rightsizing, it is possible to obtain the same changes in the *Wd* index at different values of the piston stroke and the cylinder diameter, which results from the different values of the coefficients A and B (see formula 12). The downsizing/rightsizing combinations are presented in the form of a matrix of changes in the coefficients A and B - **Figure 3**. The matrix can show two volatility zones of the

Unlike all the others, the author defined the downsizing index (*Wd*) based on the degrees of changes in the components describing the cylindrical combustion chamber (equivalent volume), which dominates the design of internal combustion engines [19]. According to this definition, the downsizing index can be described as

*<sup>S</sup> <sup>B</sup>* <sup>¼</sup> *Dd*

*<sup>D</sup>* (12)

reduction or increase of the swept volume.

*Work Cycle of Internal Combustion Engine Due to Rightsizing*

*DOI: http://dx.doi.org/10.5772/intechopen.97144*

in formula (12).

**7**

**Figure 1.**

*The idea of downsizing.*

which, taking into account the close relationship between fuel consumption and the concentration of carbon dioxide in exhaust gases, will change into (8)

$$\text{CO}\_{2d} = \text{CO}\_2\left(\frac{\text{g}\_{ed}}{\text{g}\_{\epsilon}}\right) \tag{8}$$

The changes caused by the downsizing idea can be illustrated in the diagrams - **Figure 1**.

If, in the downsizing of the internal combustion engine, the reduction of the engine speed (*downspeeding*) is made, the effect of reducing fuel consumption and limiting carbon dioxide emissions will be enhanced. For this case, assuming the stability of the parameter values as in Eq. (9)

$$V\_{sd2} = V\_{sd1}, \\ N\_{ed2} = N\_{ed1}, \\ \tau\_{d2} = \tau\_{d1}, \tag{9}$$

and changing only the speed *nd2 < nd1*, one gets (10) and (11)

$$p\_{ed1}V\_{sd1}\frac{n\_{d1}}{30\tau\_{d1}} = p\_{ed2}V\_{sd2}\frac{n\_{d2}}{30\tau\_{d2}}\tag{10}$$

$$p\_{ed2} = p\_{ed1} \left(\frac{n\_{d1}}{n\_{d2}}\right) \tag{11}$$

Index 1 represents the downsized base engine.

Index 2 denotes the downsized engine with changed (reduced) rotational speed. The measure of engine modernization, both for downsizing and rightsizing, is the degree (index) of changes, which is defined in various ways [19–21]. Regardless *Work Cycle of Internal Combustion Engine Due to Rightsizing DOI: http://dx.doi.org/10.5772/intechopen.97144*

**Figure 1.** *The idea of downsizing.*

*ped* ¼ *pe*

where the useful efficiency *η<sup>e</sup>* is expressed by the relation (5)

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

effective pressure (*pe = BMEP*).

stability of the parameter values as in Eq. (9)

Eq. (7) is obtained

**Figure 1**.

**6**

*Vss Vssd* 

In turn, fuel consumption expressed as the specific value (*ge*) can be written as (4)

*ge* <sup>¼</sup> <sup>1</sup> *ηeWu*

*<sup>η</sup><sup>e</sup>* <sup>¼</sup> *MRLppeTo ηvWupo*

With a reasonable assumption of unchanged value beyond the engine operating after downsizing, useful efficiency becomes dependent only on the brake mean

Brake specific fuel consumption (*BSFC = ge*) can also be expressed by defining

*ge* <sup>¼</sup> *Ge Ne*

Keeping the constant useful power after downsizing, i.e. *Ned = Ne*, the following

<sup>¼</sup> *Ge ge*

which, taking into account the close relationship between fuel consumption and

The changes caused by the downsizing idea can be illustrated in the diagrams -

If, in the downsizing of the internal combustion engine, the reduction of the engine speed (*downspeeding*) is made, the effect of reducing fuel consumption and limiting carbon dioxide emissions will be enhanced. For this case, assuming the

*ged ge* 

*Vssd*<sup>2</sup> ¼ *Vssd*1, *Ned*<sup>2</sup> ¼ *Ned*1, *τ<sup>d</sup>*<sup>2</sup> ¼ *τd*1, (9)

*nd*<sup>2</sup> 30*τ<sup>d</sup>*<sup>2</sup>

¼ *ped*2*Vssd*<sup>2</sup>

*nd*<sup>1</sup> *nd*<sup>2</sup> 

Index 2 denotes the downsized engine with changed (reduced) rotational speed. The measure of engine modernization, both for downsizing and rightsizing, is the degree (index) of changes, which is defined in various ways [19–21]. Regardless

the actual amount of fuel burned per unit time, giving the unit of power (6).

*Ged ged*

the concentration of carbon dioxide in exhaust gases, will change into (8)

and changing only the speed *nd2 < nd1*, one gets (10) and (11)

*nd*<sup>1</sup> 30*τ<sup>d</sup>*<sup>1</sup>

*ped*<sup>2</sup> ¼ *ped*<sup>1</sup>

*ped*1*Vssd*<sup>1</sup>

Index 1 represents the downsized base engine.

*CO*2*<sup>d</sup>* ¼ *CO*<sup>2</sup>

(3)

(4)

(5)

(6)

(7)

(8)

(10)

(11)

of the definition, this indicator shows the change or degree of residue after the reduction or increase of the swept volume.

Unlike all the others, the author defined the downsizing index (*Wd*) based on the degrees of changes in the components describing the cylindrical combustion chamber (equivalent volume), which dominates the design of internal combustion engines [19]. According to this definition, the downsizing index can be described as in formula (12).

$$W\_d = \mathbf{1} - AB^2 \qquad \text{dla} \quad A = \frac{\mathbf{S}\_d}{\mathbf{S}} \qquad B = \frac{D\_d}{D} \tag{12}$$

In the graphic interpretation, theoretically and practically three forms of changes in the swept volume can be distinguished - as in **Figure 2**.

By implementing the idea of rightsizing, it is possible to obtain the same changes in the *Wd* index at different values of the piston stroke and the cylinder diameter, which results from the different values of the coefficients A and B (see formula 12). The downsizing/rightsizing combinations are presented in the form of a matrix of changes in the coefficients A and B - **Figure 3**. The matrix can show two volatility zones of the *Wd* indicator: downsizing and upsizing, important in considering rightsizing.

Having knowledge of the design of the combustion chamber and the crank system in the commonly accepted geometric relationships between the cylinder diameter and the stroke [18, 22], as well as based on the actual relationships of these parameters determined on the basis of the engines from the Engine of the Year competition over the years 1999–2019 [19, 20] it was possible to determine the real ranges of variability of the ratio of the cylinder diameter to the piston stroke, which is from 0.77 to 1.30, which results in the value of the *Wd* index in the range of minus �1.20 on the upsizing side and plus +0.51 in the case of downsizing.

#### *Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

reduced and the emission of harmful exhaust components is significantly reduced due to the lower temperature in the combustion chamber. Additionally, a lower noise level is achieved, which significantly improves the comfort of operation [28]. On the other hand, the implementation of direct gasoline injection in spark ignition engines resulted in greater positive effects in the economic and ecological balance of engine development. The first attempts to inject gasoline directly into the combustion chamber were carried out by Jonas Hesselman in 1925, but only the solution proposed by Mitsubishi in 1996 brought success in development. This solution is known as GDI - Gasoline Direct Injection [29]. Gasoline injection, carried out in at least two phases during the intake and compression stroke, allows for stratified combustion, including combustion of very poor mixtures (50: 1 versus stoichiometric - conventional 14.7: 1), which in turn helps to increase the compression ratio without knocking effect. The use of a special combustion chamber geometry in the piston crown and thus achieving a load swirl increases engine power with a simultaneous reduction in fuel consumption. The disadvantage of this system is, unfortunately, the increase in nitrogen oxides emissions, which means that the engine must be equipped with a reducing catalyst and exhaust gas recirculation system. Of great importance in the implementation of GDI is control, including adaptive systems [30]. The use of direct injection fits very well into the architecture of the engine covered by downsizing/rightsizing because it directly complements

Another downsizing/rightsizing support system is the charging, the presence of which is essential for proper cylinder filling. As early as 1885, Gottlieb Daimler noticed the need for charging to increase the filling level in his patent about the need to increase the air pressure above atmospheric at the beginning of each cycle [18]. Then came the concept of recycling the energy wasted with the exhaust gas outlet and in 1916 Auguste Reteau built the first turbocharger. For many years, the concept of a single turbocharger functioned until the appearance of the Honeywell turbocharger, where, due to the limited response time to changes in engine load on a common axle, two compressor wheels appeared next to one turbine. The engine with such a system works more efficiently, especially in the lower engine speed (rpm) and load ranges. In the following years, various solutions began to appear, including variable VNT (Variable Nozzle Turbine) settings. An interesting solution is the system of two turbochargers working in parallel, which replace one large one. Thanks to this solution, the turbochargers are smaller (in line with the downsizing

There are also combinations of mechanical, electric and traditional charging.

The improvement of volumetric efficiency is also achieved by the application of variable valve timing systems. The variable valve timing system ensures that the angles and times of opening and closing the valves are matched to the current load

There are many variable valve timing systems which undergo successive design transformations and take different names depending on the manufacturer [34]. The first variable valve timing system appeared in 1981 on Alfa Romeo engines, but it was only the introduction of electronic control in 1989 by Honda that allowed the development of this design known as VTEC (Variable Valve Timing and lift Electronic Control), and in the latest version i- VTEC (i - intelligent system that works

In contrast, the VarioCam system, designed by Porsche in 1992, altered the position of the valves by changing the tension in the chain connecting the intake

[31–33]. Supercharging is the simplest form of supporting the downsizing/ rightsizing engine, both in terms of power loss and by creating conditions for

the power loss resulting from changes in geometry.

*Work Cycle of Internal Combustion Engine Due to Rightsizing*

*DOI: http://dx.doi.org/10.5772/intechopen.97144*

idea), which results in less heat loss to the atmosphere.

burning poor mixtures to meet ecological requirements.

and engine speed.

ahead).

**9**

**Figure 2.** *Forms of downsizing [19].*


#### **Figure 3.**

*Matrix of changes to the downsizing/rightsizing index according to various combinations of coefficients A and B (according to formula 12).*

In order to maintain the operational parameters of the internal combustion engine, while reducing its stroke volume, it is necessary to implement new or intensify the existing functions performed by individual structural and functional systems in the engine. Among them, an important place is occupied by: direct fuel injection, charging, variable valve timing, variable compression ratio. And the whole thing is controlled by electronics [23–25].

The idea of direct fuel injection developed differently in the two different engine types (diesel and gasoline). It has been used almost always in diesel engines, but the implementation of the Common Rail system by the Denso/Toyota corporation played a special role. It happened in 1995, although the idea was known as early as 1916 (Vickers company) [26]. However, at that time, there was no technology of obtaining high pressure, atomization of fuel drops and the possibility of multiple fuel injection in one cycle [27]. Today, as a result of this, fuel consumption is

#### *Work Cycle of Internal Combustion Engine Due to Rightsizing DOI: http://dx.doi.org/10.5772/intechopen.97144*

reduced and the emission of harmful exhaust components is significantly reduced due to the lower temperature in the combustion chamber. Additionally, a lower noise level is achieved, which significantly improves the comfort of operation [28].

On the other hand, the implementation of direct gasoline injection in spark ignition engines resulted in greater positive effects in the economic and ecological balance of engine development. The first attempts to inject gasoline directly into the combustion chamber were carried out by Jonas Hesselman in 1925, but only the solution proposed by Mitsubishi in 1996 brought success in development. This solution is known as GDI - Gasoline Direct Injection [29]. Gasoline injection, carried out in at least two phases during the intake and compression stroke, allows for stratified combustion, including combustion of very poor mixtures (50: 1 versus stoichiometric - conventional 14.7: 1), which in turn helps to increase the compression ratio without knocking effect. The use of a special combustion chamber geometry in the piston crown and thus achieving a load swirl increases engine power with a simultaneous reduction in fuel consumption. The disadvantage of this system is, unfortunately, the increase in nitrogen oxides emissions, which means that the engine must be equipped with a reducing catalyst and exhaust gas recirculation system. Of great importance in the implementation of GDI is control, including adaptive systems [30]. The use of direct injection fits very well into the architecture of the engine covered by downsizing/rightsizing because it directly complements the power loss resulting from changes in geometry.

Another downsizing/rightsizing support system is the charging, the presence of which is essential for proper cylinder filling. As early as 1885, Gottlieb Daimler noticed the need for charging to increase the filling level in his patent about the need to increase the air pressure above atmospheric at the beginning of each cycle [18]. Then came the concept of recycling the energy wasted with the exhaust gas outlet and in 1916 Auguste Reteau built the first turbocharger. For many years, the concept of a single turbocharger functioned until the appearance of the Honeywell turbocharger, where, due to the limited response time to changes in engine load on a common axle, two compressor wheels appeared next to one turbine. The engine with such a system works more efficiently, especially in the lower engine speed (rpm) and load ranges. In the following years, various solutions began to appear, including variable VNT (Variable Nozzle Turbine) settings. An interesting solution is the system of two turbochargers working in parallel, which replace one large one. Thanks to this solution, the turbochargers are smaller (in line with the downsizing idea), which results in less heat loss to the atmosphere.

There are also combinations of mechanical, electric and traditional charging. [31–33]. Supercharging is the simplest form of supporting the downsizing/ rightsizing engine, both in terms of power loss and by creating conditions for burning poor mixtures to meet ecological requirements.

The improvement of volumetric efficiency is also achieved by the application of variable valve timing systems. The variable valve timing system ensures that the angles and times of opening and closing the valves are matched to the current load and engine speed.

There are many variable valve timing systems which undergo successive design transformations and take different names depending on the manufacturer [34]. The first variable valve timing system appeared in 1981 on Alfa Romeo engines, but it was only the introduction of electronic control in 1989 by Honda that allowed the development of this design known as VTEC (Variable Valve Timing and lift Electronic Control), and in the latest version i- VTEC (i - intelligent system that works ahead).

In contrast, the VarioCam system, designed by Porsche in 1992, altered the position of the valves by changing the tension in the chain connecting the intake

In order to maintain the operational parameters of the internal combustion engine, while reducing its stroke volume, it is necessary to implement new or intensify the existing functions performed by individual structural and functional systems in the engine. Among them, an important place is occupied by: direct fuel injection, charging, variable valve timing, variable compression ratio. And the

*Matrix of changes to the downsizing/rightsizing index according to various combinations of coefficients A and B*

The idea of direct fuel injection developed differently in the two different engine types (diesel and gasoline). It has been used almost always in diesel engines, but the implementation of the Common Rail system by the Denso/Toyota corporation played a special role. It happened in 1995, although the idea was known as early as 1916 (Vickers company) [26]. However, at that time, there was no technology of obtaining high pressure, atomization of fuel drops and the possibility of multiple fuel injection in one cycle [27]. Today, as a result of this, fuel consumption is

whole thing is controlled by electronics [23–25].

**Figure 2.**

**Figure 3.**

**8**

*(according to formula 12).*

*Forms of downsizing [19].*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

and exhaust camshafts. Today the system is developed and also offers valve lift capability. Another example is the Valvetronic system from BMW with full control of the intake valve lift, which significantly reduces flow losses and the reaction time to load changes is reduced to a minimum.

Yet another example in this field is Ford's TI-VCT (Twin Independent - Variable Camshaft Timing) system of independent inlet and outlet valve operation, whose main advantage over other systems is better cylinder filling and scavenging the combustion chamber.

The variable valve timing system is a good complement to the downsizing/ rightsizing technique by being able to reduce flow losses due to smaller valve dimensions and by ensuring that the combustion chamber is properly filled to maintain or increase engine efficiency.

When supercharging spark-ignition engines, there may be a risk of spontaneous combustion, which is inherently undesirable. In order to prevent this, the compression ratio should be lowered, which in turn determines the pressure in the combustion chamber, and this affects the engine power throughout its entire operating range. The solution to this problem is a system with a variable compression ratio.

The principle of operation of the variable compression ratio system - VCR is associated with a change in the volume of the compression chamber with the change of load. There are several technical solutions to this issue. One of them is the change of stroke in the crank mechanism (Multi Cycle Engine 5, implemented by Peugeot).

Another way is the angular displacement of the cylinder head offered by SAAB (SVC system - Saab Variable Compression). Yet another solution is the dynamic movement of the entire crank system (Cortina VC - Variable Compression). The GoEngine solution is structurally interesting as it provides a change in the compression ratio in the range from 8: 1 to 18: 1. A significant advantage of this system is the possibility of a significant (up to 20%) extension of the expansion stroke in relation to the compression stroke, which provides better conditions for burning the fuel dose, generates more favorable pressure distribution on the piston crown and lowers the exhaust gas temperature. A variable compression ratio system, by varying the cylinder volume, can be considered one of the forms of dynamic downsizing/rightsizing, not as a support system.

From engineering practice, there are a number of examples of the development of the downsizing/rightsizing idea. We can even mention the engines installed in Ford or Volkswagen vehicles.

The Ford's engine with a displacement of 2.3 dm<sup>3</sup> V6 was reduced to 2.0 dm<sup>3</sup> and 1.6 dm<sup>3</sup> , to finally reach the spectacular 0.999 dm<sup>3</sup> EcoBoost - **Figure 4**. Some people consider the engine with a displacement of 5.0 dm3 Coyote to be the progenitor of all downsizing/rightsizing changes. This makes changes a kind of cascade of actions.

pro-ecological solutions and are included in marketing names, for example:

*The specific power of combustion engines together with the carbon dioxide emissions of the "winners" in the*

**3. Efficiency of the generalized engine work cycle in terms of**

In the combustion chamber of a reciprocating internal combustion engine, the fuel mixed with air creates a working medium that undergoes thermodynamic changes, related, among other things, to the volume of the combustion space. These changes are repeatable, although their magnitude depends on the current operating conditions of the engine. The occurring transformations create the engine work cycle, mathematically described in various ways [36–38]. In a generalized form, corresponding to all known theories of internal combustion engines, the work cycle

EcoBoost/Econetic (Ford) or Blue Motion (Volkswagen) [2].

**rightsizing - research methodology**

*engine of the year competition in all categories.*

*Rightsizing on the example of Ford's engines.*

*Work Cycle of Internal Combustion Engine Due to Rightsizing*

*DOI: http://dx.doi.org/10.5772/intechopen.97144*

**Figure 5.**

**11**

**Figure 4.**

In turn, the Volkswagen engines changed the displacement from 2.8 dm3 or 2.0 dm3 to 1.8 dm3 , and then to 1.4 dm3 , fulfilling the downsizing assumption, and with sustainable development (rightsizing) the 1.4 dm3 engine was replaced with 1.5 dm3 .

On a large scale, the trend of changing the displacement volume is well represented by the engines considered in the international competition Engine of The Year, which since 1999 has been organized by the magazine Engine Technology International - UK & International Press [35]. The winning engines in all categories show a clear trend of change in displacement over the years. It is expressed by an increase in the specific power and a decrease in carbon dioxide emissions, which increase with a decrease in the stroke volume - **Figure 5**.

In automotive practice, internal combustion engines designed in the downsizing and rightsizing technique can be found in cars with a whole package of

*Work Cycle of Internal Combustion Engine Due to Rightsizing DOI: http://dx.doi.org/10.5772/intechopen.97144*

**Figure 4.** *Rightsizing on the example of Ford's engines.*

**Figure 5.**

and exhaust camshafts. Today the system is developed and also offers valve lift capability. Another example is the Valvetronic system from BMW with full control of the intake valve lift, which significantly reduces flow losses and the reaction time

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

Yet another example in this field is Ford's TI-VCT (Twin Independent - Variable Camshaft Timing) system of independent inlet and outlet valve operation, whose main advantage over other systems is better cylinder filling and scavenging the

When supercharging spark-ignition engines, there may be a risk of spontaneous combustion, which is inherently undesirable. In order to prevent this, the compression ratio should be lowered, which in turn determines the pressure in the combustion chamber, and this affects the engine power throughout its entire operating range. The solution to this problem is a system with a variable compression ratio. The principle of operation of the variable compression ratio system - VCR is associated with a change in the volume of the compression chamber with the change of load. There are several technical solutions to this issue. One of them is the change of stroke in the crank mechanism (Multi Cycle Engine 5, implemented by

Another way is the angular displacement of the cylinder head offered by SAAB (SVC system - Saab Variable Compression). Yet another solution is the dynamic movement of the entire crank system (Cortina VC - Variable Compression). The GoEngine solution is structurally interesting as it provides a change in the compression ratio in the range from 8: 1 to 18: 1. A significant advantage of this system is the possibility of a significant (up to 20%) extension of the expansion stroke in relation to the compression stroke, which provides better conditions for burning the fuel dose, generates more favorable pressure distribution on the piston crown and lowers the exhaust gas temperature. A variable compression ratio system, by vary-

From engineering practice, there are a number of examples of the development of the downsizing/rightsizing idea. We can even mention the engines installed in

The Ford's engine with a displacement of 2.3 dm<sup>3</sup> V6 was reduced to 2.0 dm<sup>3</sup> and

, to finally reach the spectacular 0.999 dm<sup>3</sup> EcoBoost - **Figure 4**. Some people consider the engine with a displacement of 5.0 dm3 Coyote to be the progenitor of all downsizing/rightsizing changes. This makes changes a kind of cascade

In turn, the Volkswagen engines changed the displacement from 2.8 dm3 or 2.0 dm3

In automotive practice, internal combustion engines designed in the downsizing

tainable development (rightsizing) the 1.4 dm3 engine was replaced with 1.5 dm3

On a large scale, the trend of changing the displacement volume is well represented by the engines considered in the international competition Engine of The Year, which since 1999 has been organized by the magazine Engine Technology International - UK & International Press [35]. The winning engines in all categories show a clear trend of change in displacement over the years. It is expressed by an increase in the specific power and a decrease in carbon dioxide emissions, which

and rightsizing technique can be found in cars with a whole package of

, fulfilling the downsizing assumption, and with sus-

.

ing the cylinder volume, can be considered one of the forms of dynamic

downsizing/rightsizing, not as a support system.

Ford or Volkswagen vehicles.

, and then to 1.4 dm3

increase with a decrease in the stroke volume - **Figure 5**.

The variable valve timing system is a good complement to the downsizing/ rightsizing technique by being able to reduce flow losses due to smaller valve dimensions and by ensuring that the combustion chamber is properly filled to

to load changes is reduced to a minimum.

maintain or increase engine efficiency.

combustion chamber.

Peugeot).

1.6 dm<sup>3</sup>

of actions.

to 1.8 dm3

**10**

*The specific power of combustion engines together with the carbon dioxide emissions of the "winners" in the engine of the year competition in all categories.*

pro-ecological solutions and are included in marketing names, for example: EcoBoost/Econetic (Ford) or Blue Motion (Volkswagen) [2].

#### **3. Efficiency of the generalized engine work cycle in terms of rightsizing - research methodology**

In the combustion chamber of a reciprocating internal combustion engine, the fuel mixed with air creates a working medium that undergoes thermodynamic changes, related, among other things, to the volume of the combustion space. These changes are repeatable, although their magnitude depends on the current operating conditions of the engine. The occurring transformations create the engine work cycle, mathematically described in various ways [36–38]. In a generalized form, corresponding to all known theories of internal combustion engines, the work cycle

**Figure 6.** *Generalized thermodynamic work cycle of a four-stroke internal combustion engine [36].*

can be described by the efficiency (ηt) as per formula (13) and expressed graphically as in **Figure 6**.

$$\eta\_t = 1 - \frac{\frac{\lambda\_p \rho\_p \epsilon\_r^{\kappa - 1}}{\delta^{\kappa - 1}} + \kappa(\rho' - 1) - \rho'}{\varepsilon\_s^{\kappa - 1} \left\{ \lambda\_p \left[ \kappa \rho\_p - (\kappa - 1) \left( 1 + \rho\_p \ln \rho\_T \right) \right] - 1 \right\}} \tag{13}$$

The individual dimensionless quantities appearing in formula (13) are described in accordance with **Figure 6** [19].

• degree of pressure increase during isochoric heat transfer

$$
\lambda\_p = \frac{p\_{x'}}{p\_c} = \frac{p\_{x^\*}}{p\_c} \tag{14}
$$

• degree of pre-compression when heat is drained at constant pressure

*<sup>ρ</sup>*<sup>0</sup> <sup>¼</sup> *Vd Va*

*<sup>ε</sup>* <sup>¼</sup> *Vb Vc*

� �*<sup>κ</sup>*�<sup>1</sup> <sup>þ</sup> *<sup>κ</sup> Vb*

� � � ð Þ *<sup>κ</sup>* � <sup>1</sup> <sup>1</sup> <sup>þ</sup> *Vz*″

*λp, Va, Vz, Vz", κ* are components resulting from the properties of the fuel used and the logistics of the combustion process, while *Vb* and *Vc* are design parameters of the internal combustion engine related to the combustion space, and therefore

The introduction to formula (22) of the variables A and B from formula (12) gives a full picture of changes in thermodynamic transformations in the theoretical cycle of the downsizing/rightsizing engine. When assessing the effectiveness of

1.all the considered components are subject to change, that is: the displacement volume Vsd 6¼ Vs together with the compression volume *Vcd* 6¼ *Vc* and the

> 0 @

2 4

If we assume that the selection of the compression ratio for the downsizing/ rightsizing engine will be made on the basis of experimental data, e.g. by comparing the compression ratio values of the engines included in the *Engine of the Year* competition, then for typical examples the relationship between *ε* and *ε<sup>d</sup>* was

*Vs*þ*Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> � �*AB*<sup>2</sup> *Vad*

� ð Þ *<sup>κ</sup><sup>d</sup>* � <sup>1</sup> <sup>1</sup> <sup>þ</sup> *Vz*″*<sup>d</sup>*

� � " # ! � <sup>1</sup> ( )

ε<sup>d</sup> ¼ 0*:*547*ε* þ 4*:*239 (24)

1 A � 1

*Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> *AB*<sup>2</sup> !

3

<sup>5</sup> � *Vs*þ*Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup>

ln *Vzd Vz*″*<sup>d</sup>*

*<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> � �*AB*<sup>2</sup> *Vad*

(23)

By introducing the quantities expressed by the formulas (14)–(21) into the formula (13), one can obtain relationships that emphasize the changes in various volumes, which can be used to describe the changes caused by downsizing (22)

*<sup>ρ</sup><sup>T</sup>* <sup>¼</sup> *Vz Vz*″

• degree of expansion during isothermal heat transfer

*Work Cycle of Internal Combustion Engine Due to Rightsizing*

*DOI: http://dx.doi.org/10.5772/intechopen.97144*

*λp Vz*″ *Vc* � � *Va Vc* ð Þ*<sup>κ</sup>*�<sup>1</sup> *Vb Vz*

*λ<sup>p</sup> κ Vz*″ *Vc*

applying the rightsizing idea, three cases can be considered:

*Vad Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> *AB*<sup>2</sup> !*<sup>κ</sup>d*�<sup>1</sup>

1 A

*λpd κ<sup>d</sup> Vz*″*<sup>d</sup> Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> *AB*<sup>2</sup> !

*<sup>κ</sup>d*�<sup>1</sup> þ *κ<sup>d</sup>*

• geometric compression ratio

*Va Vc* � �*<sup>κ</sup>*�<sup>1</sup>

related to the rightsizing operation.

compression ratio ε<sup>d</sup> 6¼ ε (23)

*Vz*″*<sup>d</sup> Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> *AB*<sup>2</sup> !

0 @

*Vs*þ*Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> � �*AB*<sup>2</sup> *Vzd*

*λpd*

*Vad Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> *AB*<sup>2</sup> !*<sup>κ</sup>d*�<sup>1</sup>

identified [19] (16).

*η<sup>t</sup>* ¼ 1 �

*ηtd* ¼ 1 �

**13**

<sup>¼</sup> *Vb Va*

<sup>¼</sup> *Vd Vc*

*Va* � � � <sup>1</sup> h i � *Vb*

h i � � � � � <sup>1</sup>

*Va*

*Vz*″

*Vc* � � ln *Vz*

n o (22)

(19)

(20)

(21)

• degree of expansion during isobaric heat transfer

$$
\rho\_p = \frac{V\_{x^\*}}{V\_{x'}} = \frac{V\_{x^\*}}{V\_{x}} \tag{15}
$$

• effective compression ratio

$$
\varepsilon\_t = \frac{V\_a}{V\_c} \tag{16}
$$

• isentropic exponent

$$\kappa = \frac{c\_p}{c\_v} \tag{17}$$

• degree of another expansion process

$$
\delta = \frac{V\_b}{V\_x} \tag{18}
$$

*Work Cycle of Internal Combustion Engine Due to Rightsizing DOI: http://dx.doi.org/10.5772/intechopen.97144*

• degree of pre-compression when heat is drained at constant pressure

$$
\rho' = \frac{V\_d}{V\_a} = \frac{V\_b}{V\_a} \tag{19}
$$

• degree of expansion during isothermal heat transfer

$$
\rho\_T = \frac{V\_x}{V\_{x^\*}} \tag{20}
$$

• geometric compression ratio

$$
\varepsilon = \frac{V\_b}{V\_c} = \frac{V\_d}{V\_c} \tag{21}
$$

By introducing the quantities expressed by the formulas (14)–(21) into the formula (13), one can obtain relationships that emphasize the changes in various volumes, which can be used to describe the changes caused by downsizing (22)

$$\eta\_{t} = \mathbf{1} - \frac{\frac{\lambda\_{p} \left(\frac{V\_{r}}{V\_{c}}\right) \left(\frac{V\_{d}}{V\_{c}}\right)^{\kappa - 1}}{\left(\frac{V\_{d}}{V\_{c}}\right)^{\kappa - 1}} + \kappa \left[\left(\frac{V\_{b}}{V\_{c}}\right) - \mathbf{1}\right] - \frac{V\_{b}}{V\_{s}}} \\ \tag{22}$$
  $\eta\_{t} = \mathbf{1} - \frac{\left(\frac{V\_{d}}{V\_{c}}\right)^{\kappa - 1} \left\{\lambda\_{p} \left[\kappa \left(\frac{V\_{r}}{V\_{c}}\right) - (\kappa - \mathbf{1}) \left(\mathbf{1} + \left(\frac{V\_{r}}{V\_{c}}\right) \ln\left(\frac{V\_{r}}{V\_{c'}}\right)\right)\right] - \mathbf{1}\right\}}{\left(\frac{V\_{d}}{V\_{c}}\right)^{\kappa - 1} \left(\frac{V\_{d}}{V\_{c}}\right) \left[\lambda\_{p} - \mathbf{1}\right] - \mathbf{1}\right)} \tag{23}$ 

*λp, Va, Vz, Vz", κ* are components resulting from the properties of the fuel used and the logistics of the combustion process, while *Vb* and *Vc* are design parameters of the internal combustion engine related to the combustion space, and therefore related to the rightsizing operation.

The introduction to formula (22) of the variables A and B from formula (12) gives a full picture of changes in thermodynamic transformations in the theoretical cycle of the downsizing/rightsizing engine. When assessing the effectiveness of applying the rightsizing idea, three cases can be considered:

1.all the considered components are subject to change, that is: the displacement volume Vsd 6¼ Vs together with the compression volume *Vcd* 6¼ *Vc* and the compression ratio ε<sup>d</sup> 6¼ ε (23)

*ηtd* ¼ 1 � *λpd Vz*″*<sup>d</sup> Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> *AB*<sup>2</sup> ! *Vad Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> *AB*<sup>2</sup> !*<sup>κ</sup>d*�<sup>1</sup> *Vs*þ*Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> � �*AB*<sup>2</sup> *Vzd* 0 @ 1 A *<sup>κ</sup>d*�<sup>1</sup> þ *κ<sup>d</sup> Vs*þ*Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> � �*AB*<sup>2</sup> *Vad* 0 @ 1 A � 1 2 4 3 <sup>5</sup> � *Vs*þ*Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> � �*AB*<sup>2</sup> *Vad Vad Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> *AB*<sup>2</sup> !*<sup>κ</sup>d*�<sup>1</sup> *λpd κ<sup>d</sup> Vz*″*<sup>d</sup> Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> *AB*<sup>2</sup> ! � ð Þ *<sup>κ</sup><sup>d</sup>* � <sup>1</sup> <sup>1</sup> <sup>þ</sup> *Vz*″*<sup>d</sup> Vc* ð Þ *<sup>ε</sup>*�<sup>1</sup> *<sup>ε</sup><sup>d</sup>* ð Þ �<sup>1</sup> *AB*<sup>2</sup> ! ln *Vzd Vz*″*<sup>d</sup>* � � " # ! � <sup>1</sup> ( ) (23)

If we assume that the selection of the compression ratio for the downsizing/ rightsizing engine will be made on the basis of experimental data, e.g. by comparing the compression ratio values of the engines included in the *Engine of the Year* competition, then for typical examples the relationship between *ε* and *ε<sup>d</sup>* was identified [19] (16).

$$
\varepsilon\_{\rm d} = 0.547 \varepsilon + 4.239 \tag{24}
$$

can be described by the efficiency (ηt) as per formula (13) and expressed

*Generalized thermodynamic work cycle of a four-stroke internal combustion engine [36].*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

*λpρpεκ*�<sup>1</sup> *<sup>s</sup>*

*<sup>λ</sup><sup>p</sup>* <sup>¼</sup> *pz*<sup>0</sup> *pc*

*<sup>ρ</sup><sup>p</sup>* <sup>¼</sup> *Vz*″ *Vz*<sup>0</sup>

> *<sup>ε</sup><sup>s</sup>* <sup>¼</sup> *Va Vc*

*<sup>κ</sup>* <sup>¼</sup> *cp cv*

*<sup>δ</sup>* <sup>¼</sup> *Vb Vz*

• degree of pressure increase during isochoric heat transfer

• degree of expansion during isobaric heat transfer

*εκ*�<sup>1</sup> *<sup>s</sup> λ<sup>p</sup> κρ<sup>p</sup>* � ð Þ *κ* � 1 1 þ *ρ<sup>p</sup>* ln *ρ<sup>T</sup>*

The individual dimensionless quantities appearing in formula (13) are described

<sup>¼</sup> *pz*″ *pc*

<sup>¼</sup> *Vz*″ *Vc*

*δκ*�<sup>1</sup> þ *κ ρ*ð Þ� <sup>0</sup> � 1 *ρ*<sup>0</sup>

h i � �

n o (13)

� 1

(14)

(15)

(16)

(17)

(18)

graphically as in **Figure 6**.

**Figure 6.**

*η<sup>t</sup>* ¼ 1 �

in accordance with **Figure 6** [19].

• effective compression ratio

• degree of another expansion process

• isentropic exponent

**12**

#### *Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

It means the possibility of introducing a new coefficient (C), expressed by the relation (25).

$$\frac{\varepsilon\_d - 1}{\varepsilon - 1} = \mathcal{C} \tag{25}$$

After taking into account the dependence (25), the formula describing the theoretical efficiency of the engine work cycle takes the form (26).

$$\eta\_{td} = \mathbf{1} - \frac{\frac{\lambda\_{pd}\left(\frac{V\_{r'd}}{V\_{r}\Delta^2}\right)\left(\frac{V\_{ad}}{V\_{r}\Delta^2}\right)^{\kappa\_d - 1}}{\left(\frac{\left(V\_{r} + V\_{c}\frac{1}{C}\right)AB\right)^{\kappa\_d - 1}} + \kappa\_d\left[\left(\frac{\left(V\_{r} + V\_{c}\frac{1}{C}\right)AB^2}{V\_{ad}}\right) - \mathbf{1}\right] - \frac{\left(V\_{r} + V\_{c}\frac{1}{C}\right)AB^2}{V\_{ad}}}}{\left(\frac{V\_{ad}}{V\_{c}\Delta^2}\right)^{\kappa\_d - 1}\left\{\lambda\_{pd}\left[\kappa\_d\left(\frac{V\_{r'd}}{V\_{c}\frac{1}{C}}\right) - (\kappa\_d - 1)\left(\mathbf{1} + \left(\frac{V\_{r'd}}{V\_{c'}\frac{4\Delta^2}{C}}\right)\ln\left(\frac{V\_{d}}{V\_{c'd}}\right)\right)\right] - \mathbf{1}\right\}}\tag{26}$$

2. the following are subject to change: the swept volume *Vsd* 6¼ *Vs* and the compression volume *Vcd* 6¼ *Vc* without changing the compression ratio *ε<sup>d</sup> = ε* (27)

$$\eta\_{td} = \mathbf{1} - \frac{\frac{\lambda\_{pd} \left(\frac{V\_{x'd}}{V\_{t}AB^2}\right) \left(\frac{V\_{xd}}{V\_{t}AB^2}\right)^{\kappa\_d - 1}}{\left(\frac{(V\_{t} + V\_{t})AB^2}{V\_{xd}}\right)^{\kappa\_d - 1}} + \kappa\_d \left[\left(\frac{(V\_{t} + V\_{t})AB^2}{V\_{xd}}\right) - \mathbf{1}\right] - \frac{(V\_{t} + V\_{t})AB^2}{V\_{xd}}}{\left(\frac{V\_{xd}}{V\_{t}AB^2}\right)^{\kappa\_d - 1} \left\{\lambda\_{pd} \left[\kappa\_d \left(\frac{V\_{x'd}}{V\_{t}AB^2}\right) - (\kappa\_d - 1)\left(\mathbf{1} + \left(\frac{V\_{x'd}}{V\_{t}AB^2}\right) \ln\left(\frac{V\_{xd}}{V\_{x'd}}\right)\right)\right] - \mathbf{1}\right\}} \tag{27}$$

3. the third case is the change of the swept volume *Vsd* 6¼ *Vs* and the compression ratio *ε<sup>d</sup>* 6¼ *ε* without changing the compression space *Vcd = Vc* (28)

$$\eta\_{ld} = \mathbf{1} - \frac{\frac{\lambda\_{pd} \left(\frac{V\_{\nu d}}{V\_c}\right) \left(\frac{V\_{ad}}{V\_c}\right)^{\kappa\_d - 1}}{\left(\frac{V\_{ad} \oplus V\_c}{V\_{ad}}\right)^{\kappa\_d - 1}} + \kappa\_d \left[\left(\frac{V\_r AB^2 + V\_c}{V\_{ad}}\right) - \mathbf{1}\right] - \frac{V\_r AB^2 + V\_c}{V\_{ad}}}$$

$$\eta\_{ld} = \mathbf{1} - \frac{\left(\frac{V\_{ad}}{V\_c}\right)^{\kappa\_d - 1} \left\{\lambda\_{pd} \left[\kappa\_d \left(\frac{V\_{\nu d}}{V\_c}\right) - (\kappa\_d - 1)\left(\mathbf{1} + \left(\frac{V\_{\nu d}}{V\_c}\right) \ln\left(\frac{V\_{\nu d}}{V\_{\nu d}}\right)\right)\right] - \mathbf{1}\right\}}{\left(\mathbf{1} + \left(\frac{V\_{\nu d}}{V\_c}\right) \ln\left(\frac{V\_{\nu d}}{V\_c}\right)\right) - \mathbf{1}\right)}\tag{28}$$

The downsizing/rightsizing indexes according to formula (12) for the cascade of

For each case, apart from the factory version, theoretical changes related to the behavior of the Wd index with different coefficients A and B, taken from the matrix

In this way, a package of variables was obtained and analyzed - **Table 1**. It is worth noting that in the case of the 1.4 dm<sup>3</sup> engine, in which downsizing/ rightsizing according to the form "cylinder version" (A = 1) was intended, the rule of mutual relation of the diameter and stroke of the piston, which should be in the range (0.77–1.30) - as discussed above. Hence the decision to change the relation to

changes of the swept volume are as follows:

*Tested vehicle with a 1.8 T engine on a chassis dynamometer.*

• 2.0 dm<sup>3</sup> na 1.8 dm3 Wd = 0.09

**Figure 7.**

**Figure 8.**

*VW 1.4 dm3 engine on the test stand [39].*

*Work Cycle of Internal Combustion Engine Due to Rightsizing*

*DOI: http://dx.doi.org/10.5772/intechopen.97144*

• 2.0 dm<sup>3</sup> na 1.5 dm3 Wd = 0.25

• 2.0 dm<sup>3</sup> na 1.4 dm<sup>3</sup> Wd = 0.29

of changes - **Figure 2**, were considered.

the closest unity to A = 0.97.

**15**

In the test evaluation methodology, the real values of A and B coefficient pairs are introduced from the matrix described in **Figure 2**. This way, changes in the thermodynamic work cycle efficiency can be calculated. The rest of the data was taken from research on the 1.4 TSI, 1.5 TFSI, 1.8 T and 2.0 TDI engines, which are an example of a link in the downsizing/rightsizing chain of Volkswagen engines.

The study covered an extreme case of changes, i.e. changes in both the swept volume, compression and compression ratio (formula 23).

To evaluate the research problem, theoretical and experimental data from the tests of the VW 1.4 TSI internal combustion engine carried out at the Department of Vehicle Engineering of the Wroclaw University of Science and Technology - **Figure 7**. The next data were used from the tests on the chassis dynamometer of vehicles equipped with 1.8 T and 2.0 TDi engines - **Figure 8**.

The research data constituting the boundary conditions for the evaluation of the 1.5 TFSI engine were obtained from the literature [40].

*Work Cycle of Internal Combustion Engine Due to Rightsizing DOI: http://dx.doi.org/10.5772/intechopen.97144*

**Figure 7.** *VW 1.4 dm3 engine on the test stand [39].*

#### **Figure 8.**

It means the possibility of introducing a new coefficient (C), expressed by the

*<sup>ε</sup>* � <sup>1</sup> <sup>¼</sup> *<sup>C</sup>* (25)

� 1

*Vc AB*2 *C* � �

� 1

*VcAB*<sup>2</sup> � �

� *VsAB*<sup>2</sup>

ln *Vzd Vz*00*<sup>d</sup>*

þ*Vc Vad*

� 1

� *Vs*þ*Vc* <sup>1</sup> ð Þ*<sup>C</sup> AB*<sup>2</sup> *Vad*

ln *Vzd Vz*″*<sup>d</sup>*

� ð Þ *Vs*þ*Vc AB*<sup>2</sup> *Vad*

> ln *Vzd Vz*00*<sup>d</sup>*

� 1

� 1

(27)

(26)

*ε<sup>d</sup>* � 1

After taking into account the dependence (25), the formula describing the

*Vs*þ*Vc* <sup>1</sup> ð Þ*<sup>C</sup> AB*<sup>2</sup> *Vad* � �

� �

� ð Þ *<sup>κ</sup><sup>d</sup>* � <sup>1</sup> <sup>1</sup> <sup>þ</sup> *Vz*″*<sup>d</sup>*

*Vad* � �

� ð Þ *<sup>κ</sup><sup>d</sup>* � <sup>1</sup> <sup>1</sup> <sup>þ</sup> *Vz*00*<sup>d</sup>*

h i � � � �

n o

� 1

n o (28)

*Vc* � �

h i

� � � � � �

� �

theoretical efficiency of the engine work cycle takes the form (26).

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

*Vc AB*2 *C* � �*<sup>κ</sup>d*�<sup>1</sup>

*λpd κ<sup>d</sup>*

� �*<sup>κ</sup>d*�<sup>1</sup> <sup>þ</sup> *<sup>κ</sup><sup>d</sup>*

*VcAB*<sup>2</sup> � �*<sup>κ</sup>d*�<sup>1</sup>

ð Þ *Vs*þ*Vc AB*<sup>2</sup> *Vzd*

*λpd κ<sup>d</sup>*

*Vz*″*<sup>d</sup> Vc AB*2 *C* � �

2. the following are subject to change: the swept volume *Vsd* 6¼ *Vs* and the compression volume *Vcd* 6¼ *Vc* without changing the compression ratio *ε<sup>d</sup> = ε*

� �*<sup>κ</sup>d*�<sup>1</sup> <sup>þ</sup> *<sup>κ</sup><sup>d</sup>* ð Þ *Vs*þ*Vc AB*<sup>2</sup>

ratio *ε<sup>d</sup>* 6¼ *ε* without changing the compression space *Vcd = Vc* (28)

3. the third case is the change of the swept volume *Vsd* 6¼ *Vs* and the compression

*VsAB*<sup>2</sup> þ*Vc Vad* � �

� ð Þ *<sup>κ</sup><sup>d</sup>* � <sup>1</sup> <sup>1</sup> <sup>þ</sup> *Vz*00*<sup>d</sup>*

In the test evaluation methodology, the real values of A and B coefficient pairs are introduced from the matrix described in **Figure 2**. This way, changes in the thermodynamic work cycle efficiency can be calculated. The rest of the data was taken from research on the 1.4 TSI, 1.5 TFSI, 1.8 T and 2.0 TDI engines, which are an example of a link in the downsizing/rightsizing chain of Volkswagen engines. The study covered an extreme case of changes, i.e. changes in both the swept

To evaluate the research problem, theoretical and experimental data from the tests of the VW 1.4 TSI internal combustion engine carried out at the Department of Vehicle Engineering of the Wroclaw University of Science and Technology - **Figure 7**. The next data were used from the tests on the chassis dynamometer of

The research data constituting the boundary conditions for the evaluation of the

h i � � � �

h i

*Vz*00*<sup>d</sup> VcAB*<sup>2</sup> � �

*Vs*þ*Vc* <sup>1</sup> ð Þ*<sup>C</sup> AB*<sup>2</sup> *Vzd*

relation (25).

*ηtd* ¼ 1 �

(27)

*ηtd* ¼ 1 �

*ηtd* ¼ 1 �

**14**

*λpd Vz*″*<sup>d</sup> Vc AB*2 *C* � � *Vad*

> *λpd Vz*00*<sup>d</sup> VcAB*<sup>2</sup> � � *Vad*

*λpd Vz*00*d Vc* � � *Vad Vc* � �*<sup>κ</sup>d*�<sup>1</sup>

*Vad Vc* � �*<sup>κ</sup>d*�<sup>1</sup> *VsAB*2þ*Vc Vzd*

*λpd κ<sup>d</sup>*

� �*<sup>κ</sup>d*�<sup>1</sup> <sup>þ</sup> *<sup>κ</sup><sup>d</sup>*

volume, compression and compression ratio (formula 23).

vehicles equipped with 1.8 T and 2.0 TDi engines - **Figure 8**.

1.5 TFSI engine were obtained from the literature [40].

*Vz*00*<sup>d</sup> Vc* � �

*Vad VcAB*<sup>2</sup> � �*<sup>κ</sup>d*�<sup>1</sup>

*Vad Vc AB*2 *C* � �*<sup>κ</sup>d*�<sup>1</sup>

*Tested vehicle with a 1.8 T engine on a chassis dynamometer.*

The downsizing/rightsizing indexes according to formula (12) for the cascade of changes of the swept volume are as follows:


For each case, apart from the factory version, theoretical changes related to the behavior of the Wd index with different coefficients A and B, taken from the matrix of changes - **Figure 2**, were considered.

In this way, a package of variables was obtained and analyzed - **Table 1**.

It is worth noting that in the case of the 1.4 dm<sup>3</sup> engine, in which downsizing/ rightsizing according to the form "cylinder version" (A = 1) was intended, the rule of mutual relation of the diameter and stroke of the piston, which should be in the range (0.77–1.30) - as discussed above. Hence the decision to change the relation to the closest unity to A = 0.97.


#### **Table 1.**

*The stroke and cylinder diameter values and the corresponding coefficients define the downsizing/rightsizing area.*

The values of the A and B coefficients taken for the assessment filled successive values of the downsizing/rightsizing Wd index, ensuring their invariability within a given cylinder volume. The remaining data, filling the form of the formula for the efficiency of the comparative cycle with the equivalent volume (formula 23) and enabling the evaluation of the engine performance indicators, were obtained from the above-mentioned laboratory tests.

#### **4. Discussion of the results**

Typical operating indicators of engine work were assessed together with parameters of the thermodynamic cycle, including the efficiency of the generalized work cycle. The obtained data are presented in the form of relative changes, i.e. as a percentage of the data for the base engine 2.0 dm<sup>3</sup> - **Tables 2**–**4**.

The data contained in **Table 2** refer to the 1.8 dm<sup>3</sup> engine and confirm the correctness of the downsizing idea due to the reduction in fuel consumption by an average of 5%. Thanks to the support systems with supercharging at the forefront and control of the combustion process, even an increase in power of nearly 14% compared to the 2.0 dm3 unit was achieved.

Greater efficiency was obtained both on the theoretical and useful side. The differences between the efficiency changes *η<sup>e</sup>* and *η<sup>t</sup>* are due to exhaust losses and cooling.

It is worth emphasizing that the change of the coefficients A and B, in such a way that the downsizing index *Wd* is maintained, did not cause significant differences in the values of all examined parameters and fell within the limits of statistical significance.

The approx. 27% increase in volumetric efficiency is due to the boost system and variable valve timing set. The engine power was retained with the seemingly rea-

increase in BMEP. In the group of tested engines, it is the only engine in which the *downspeeding* concept was applied, changing the maximum value of the engine speed from 6000 to 5000 rpm. There was no significant increase in temperature in the maximum work cycle. Keeping the downsizing/rightsizing index *Wd* at the level

*Values of selected engine operating parameters 1.5 dm3 in relation to 2.0 dm3 at different values of the*

**Parameter Manufacturer-2.0 Manufacturer-1.8\_1 Test-1.8\_2 Test-1.8\_3**

*Work Cycle of Internal Combustion Engine Due to Rightsizing*

*DOI: http://dx.doi.org/10.5772/intechopen.97144*

ε 10.5 4.7 4.7 4.7 rpm 6000 8.3 8.3 8.3 n1 1.35 1.5 +0.4 1.0 n2 1.19 0 0.2 0 Tmax, K 2706 0.7 +0.1 0.5 η<sup>v</sup> 0.92 +29.1 +28.9 +29.1 BMEP 1.11 +36.8 +35.9 +36.0 BSFC, g/kWh 264 5.6 5.2 5.0 Ne, kW 110 +13.6 +13.7 +13.7 η<sup>e</sup> 0.32 +6.0 +5.5 +5.3 η<sup>t</sup> 0.45 +0.7 +0.4 +0.4

*Values of selected engine operating parameters 1.8 dm3 in relation to 2.0 dm3 at different values of the*

**Parameter Manufacturer-2.0 Manufacturer-1.5\_1 Test-1.5\_2 Test-1.5\_3**

ε 10.5 +19.1 +19.1 +19.1 rpm 6000 16.7 16.7 16.7 n1 1.35 +9.0 +9.0 +9.7 n2 1.19 2.5 2.5 2.5 Tmax, K 2706 +6.6 +6.6 +7.5 η<sup>v</sup> 0.92 +26.6 +26.6 +29.7 BMEP 1.11 +54.3 +56.7 +59.4 BSFC, g/kWh 264 19.7 20.9 19.5 Ne, kW 110 +2.9 +2.1 +0.22 η<sup>e</sup> 0.32 +24.5 +26.4 +24.2 η<sup>t</sup> 0.45 +6.9 +7.2 +6.8

**% 1.8/2.0 % 1.8/2.0 % 1.8/2.0**

**% 1.5/2.0 % 1.5/2.0 % 1.5/2.0**

of 0.25, it was shown that the change of coefficients A and B does not cause

, but less than 1.4dm<sup>3</sup>

sonable boost, which resulted in a greater than the 1.8dm<sup>3</sup>

*downsizing/rightsizing coefficients A and B (Table 1).*

**Table 2.**

**Table 3.**

**17**

*downsizing/rightsizing coefficients A and B (Table 1).*

differentiation of the theoretical work cycle efficiency.

The data contained in **Table 3** refer to the 1.5 dm<sup>3</sup> engine and confirm the correctness of the downsizing concept due to the reduction in fuel consumption by an average of nearly 20%. The proposal to reduce the stroke volume by about 25% is close to aggressive downsizing.


#### *Work Cycle of Internal Combustion Engine Due to Rightsizing DOI: http://dx.doi.org/10.5772/intechopen.97144*

#### **Table 2.**

The values of the A and B coefficients taken for the assessment filled successive values of the downsizing/rightsizing Wd index, ensuring their invariability within a given cylinder volume. The remaining data, filling the form of the formula for the efficiency of the comparative cycle with the equivalent volume (formula 23) and enabling the evaluation of the engine performance indicators, were obtained from

*The stroke and cylinder diameter values and the corresponding coefficients define the downsizing/rightsizing*

0.77 ≤ D/S ≤ 1.30

**Engines S A Sd DBDd Wd D/S Dd/Sd Remarks**

2.0 92.80 — — 82.50 — — 0 0.89 — base 1.8 0.91 84.10 1.00 82.50 0.090 0.98 Factory-1.8\_1 1.8 1.00 92.80 0.955 78.80 0.088 0.85 Test-1.8\_2 1.8 0.97 90.00 0.97 80.00 0.087 0.89 Test-1.8\_3 1.5 0.93 85.9 0.90 74.50 0.247 0.87 Factory-1.5\_1 1.5 0.75 69.60 1.00 82.5 0.250 1.19 Test-1.5\_2 1.5 1.00 92.80 0.87 71.50 0.243 0.77 Test-1.5\_3 1.4 0.82 75.60 0.93 76.50 0.291 1.01 Factory-1.4\_1 1.4 1.00 92.80 0.84 69.30 0.294 0.75 Does not meet

1.4 0.97 90.00 0.855 70.50 0.291 0.78 Test-1.4\_2 1.4 0.71 65.00 1.00 82.50 0.290 1.25 Test-1.4-3

dm<sup>3</sup> mm — mm mm — mm ———

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

Typical operating indicators of engine work were assessed together with parameters of the thermodynamic cycle, including the efficiency of the generalized work cycle. The obtained data are presented in the form of relative changes, i.e. as a

The data contained in **Table 2** refer to the 1.8 dm<sup>3</sup> engine and confirm the correctness of the downsizing idea due to the reduction in fuel consumption by an average of 5%. Thanks to the support systems with supercharging at the forefront and control of the combustion process, even an increase in power of nearly 14%

Greater efficiency was obtained both on the theoretical and useful side. The differences between the efficiency changes *η<sup>e</sup>* and *η<sup>t</sup>* are due to exhaust losses and

It is worth emphasizing that the change of the coefficients A and B, in such a way that the downsizing index *Wd* is maintained, did not cause significant differences in the values of all examined parameters and fell within the limits of statistical

The data contained in **Table 3** refer to the 1.5 dm<sup>3</sup> engine and confirm the correctness of the downsizing concept due to the reduction in fuel consumption by an average of nearly 20%. The proposal to reduce the stroke volume by about 25% is

percentage of the data for the base engine 2.0 dm<sup>3</sup> - **Tables 2**–**4**.

the above-mentioned laboratory tests.

compared to the 2.0 dm3 unit was achieved.

**4. Discussion of the results**

cooling.

**16**

**Table 1.**

*area.*

significance.

close to aggressive downsizing.

*Values of selected engine operating parameters 1.8 dm3 in relation to 2.0 dm3 at different values of the downsizing/rightsizing coefficients A and B (Table 1).*


#### **Table 3.**

*Values of selected engine operating parameters 1.5 dm3 in relation to 2.0 dm3 at different values of the downsizing/rightsizing coefficients A and B (Table 1).*

The approx. 27% increase in volumetric efficiency is due to the boost system and variable valve timing set. The engine power was retained with the seemingly reasonable boost, which resulted in a greater than the 1.8dm<sup>3</sup> , but less than 1.4dm<sup>3</sup> increase in BMEP. In the group of tested engines, it is the only engine in which the *downspeeding* concept was applied, changing the maximum value of the engine speed from 6000 to 5000 rpm. There was no significant increase in temperature in the maximum work cycle. Keeping the downsizing/rightsizing index *Wd* at the level of 0.25, it was shown that the change of coefficients A and B does not cause differentiation of the theoretical work cycle efficiency.



**Figure 9** presents the relationship between the cycle efficiency and the downsizing index, which shows that the reduction of the swept volume will be effective up to a certain limit. For the analyzed case, the implementation of an engine with a volume of 1.5 dm3 instead of 1.4 dm<sup>3</sup> is an example of this.

The problem of changing the displacement of the internal combustion engine is

A research problem was defined, which is the assessment of the influence of the differences in geometrical changes of the piston stroke and the cylinder diameter, while maintaining the same value of the downsizing index, on the efficiency of the

To fulfill the aim of the research, the relationship describing the theoretical efficiency of the general reference cycle was modified, in which instead of the stroke volume, substitute coefficients defining changes in the value of the piston's

A spectacular case of changes was adopted for the considerations, assuming a change in the swept volume and the accompanying changes in the compression space and compression ratio. The necessary data for the analysis was obtained from laboratory tests and from the literature. There have been repeatedly estimated intermediate quantities defining the efficiency of the engine supported by the

The analysis of the results shows that the efficiency of the internal combustion engine cycle is stable, regardless of the A and B coefficients, which determine the

The presence of the limit value of the downsizing/rightsizing index was also demonstrated, at which the highest level of positive change in the circulation efficiency is achieved, corresponding to the sustainability requirements.

In the next steps, the research work must be targeted at detailed on-road exhaust gas toxicity testing of downsizing/rightsizing engines. Due to the significant load on the engine structure, it will also be important to pay attention to the issues of material engineering and tribological processes related to the downsizing/

The works were carried out in the GEO-3EM research complex of the Wroclaw University of Science and Technology, in the laboratories of the Department of

This research was funded by Wroclaw University of Science and Technology,

known as downsizing, but recently it has been undergoing a transformation towards rightsizing. It is the result of a new approach to the design and operation process, which assumes balancing the customer's requirements and the manufacturer's capabilities, all in a specific environment, e.g. constantly tightening ecological standards. Anyway, the ecological aspect is the most desirable criterion for assessing the rightsizing concept, which is expressed in the pursuit of reducing fuel consumption and the resulting reduction in carbon dioxide emissions, all for the

correct selection of the engine's displacement volume.

*Work Cycle of Internal Combustion Engine Due to Rightsizing*

*DOI: http://dx.doi.org/10.5772/intechopen.97144*

stroke (A) and the cylinder diameter (B) were revealed.

geometric changes of the engine displacement volume.

obtained engine operation parameters.

**5. Summary**

engine work cycle.

rightsizing concept.

**Acknowledgements**

Automotive Engineering.

**19**

grant number MPK 9100560000/8201003902.

#### **Table 4.**

*Values of selected engine operating parameters 1.4 dm3 in relation to 2.0 dm3 at different values of the downsizing/rightsizing coefficients A and B (Table 1).*

The data in **Table 4** refer to the 1.4 dm<sup>3</sup> engine and indicate an aggressive downsizing of up to 30%. The expected effect was achieved, i.e. the specific fuel consumption was reduced by an average of 13%, which obviously translates into a reduction in carbon dioxide emissions to the atmosphere. The implementation of support systems for geometric changes resulted in a significant increase in BMEP by over 60%, which may result in a reduction in the durability of engine parts, especially in the area of the piston and crank system.

The differences between the efficiency changes *η<sup>e</sup>* and *η<sup>t</sup>* are due to losses in the exhaust and cooling systems.

The change of coefficients A and B does not significantly affect, and even the differences in values are insignificant, on the tested parameters.

From the point of view of rightsizing, it should be noted a clear relationship between the cycle efficiency and the necessary change in the stroke volume, i.e. one that will correspond to a sustainable approach to design by meeting customer needs and at the same time fulfilling the manufacturer's capabilities.

**Figure 9.**

**Figure 9** presents the relationship between the cycle efficiency and the downsizing index, which shows that the reduction of the swept volume will be effective up to a certain limit. For the analyzed case, the implementation of an engine with a volume of 1.5 dm3 instead of 1.4 dm<sup>3</sup> is an example of this.

#### **5. Summary**

The data in **Table 4** refer to the 1.4 dm<sup>3</sup> engine and indicate an aggressive downsizing of up to 30%. The expected effect was achieved, i.e. the specific fuel consumption was reduced by an average of 13%, which obviously translates into a reduction in carbon dioxide emissions to the atmosphere. The implementation of support systems for geometric changes resulted in a significant increase in BMEP by over 60%, which may result in a reduction in the durability of engine parts, espe-

*Values of selected engine operating parameters 1.4 dm3 in relation to 2.0 dm3 at different values of the*

**Parameter Manufacturer-2.0 Manufacturer-1.4\_1 Test-1.4\_2 Test-1.4\_3**

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

ε 10.5 4.7 4.7 4.7 rpm 6000 0 0 0 n1 1.35 +7.5 +7.4 +9.6 n2 1.19 3.4 3.4 2.5 Tmax, K 2706 +5.0 +5.0 +6.5 η<sup>v</sup> 0.92 +38.4 +39.0 +37.8 BMEP 1.11 +62.2 +60.4 +62.0 BSFC, g/kWh 264 13.7 12.4 14.0 Ne, kW 110 +13.6 +13.6 +13.5 η<sup>e</sup> 0.32 +15.9 +14.2 +13.5 η<sup>t</sup> 0.45 +2.7 +2.7 +2.6

**% 1.4/2.0 % 1.4/2.0 % 1.4/2.0**

The differences between the efficiency changes *η<sup>e</sup>* and *η<sup>t</sup>* are due to losses in the

The change of coefficients A and B does not significantly affect, and even the

From the point of view of rightsizing, it should be noted a clear relationship between the cycle efficiency and the necessary change in the stroke volume, i.e. one that will correspond to a sustainable approach to design by meeting customer needs

cially in the area of the piston and crank system.

*downsizing/rightsizing coefficients A and B (Table 1).*

differences in values are insignificant, on the tested parameters.

and at the same time fulfilling the manufacturer's capabilities.

*Changes in the efficiency of the engine work cycle in relation to the downsizing index.*

exhaust and cooling systems.

**Table 4.**

**Figure 9.**

**18**

The problem of changing the displacement of the internal combustion engine is known as downsizing, but recently it has been undergoing a transformation towards rightsizing. It is the result of a new approach to the design and operation process, which assumes balancing the customer's requirements and the manufacturer's capabilities, all in a specific environment, e.g. constantly tightening ecological standards. Anyway, the ecological aspect is the most desirable criterion for assessing the rightsizing concept, which is expressed in the pursuit of reducing fuel consumption and the resulting reduction in carbon dioxide emissions, all for the correct selection of the engine's displacement volume.

A research problem was defined, which is the assessment of the influence of the differences in geometrical changes of the piston stroke and the cylinder diameter, while maintaining the same value of the downsizing index, on the efficiency of the engine work cycle.

To fulfill the aim of the research, the relationship describing the theoretical efficiency of the general reference cycle was modified, in which instead of the stroke volume, substitute coefficients defining changes in the value of the piston's stroke (A) and the cylinder diameter (B) were revealed.

A spectacular case of changes was adopted for the considerations, assuming a change in the swept volume and the accompanying changes in the compression space and compression ratio. The necessary data for the analysis was obtained from laboratory tests and from the literature. There have been repeatedly estimated intermediate quantities defining the efficiency of the engine supported by the obtained engine operation parameters.

The analysis of the results shows that the efficiency of the internal combustion engine cycle is stable, regardless of the A and B coefficients, which determine the geometric changes of the engine displacement volume.

The presence of the limit value of the downsizing/rightsizing index was also demonstrated, at which the highest level of positive change in the circulation efficiency is achieved, corresponding to the sustainability requirements.

In the next steps, the research work must be targeted at detailed on-road exhaust gas toxicity testing of downsizing/rightsizing engines. Due to the significant load on the engine structure, it will also be important to pay attention to the issues of material engineering and tribological processes related to the downsizing/ rightsizing concept.

#### **Acknowledgements**

The works were carried out in the GEO-3EM research complex of the Wroclaw University of Science and Technology, in the laboratories of the Department of Automotive Engineering.

This research was funded by Wroclaw University of Science and Technology, grant number MPK 9100560000/8201003902.

**References**

[1] Warnecke W., Lueke W., Clarke L., Louis J., Kempsel S., Fuels of the Future. Proceedings of 27th International Vienna Motor Symposium, Vienna 2006.

*DOI: http://dx.doi.org/10.5772/intechopen.97144*

*Work Cycle of Internal Combustion Engine Due to Rightsizing*

[9] Song D., Hycet e-Chuang, Great Wall Motor, Hebei, China; W. Happenhofer, Great Wall Motor, Hebei, China: 1.5T High Thermal Efficiency Modular Engine Platform, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

[10] Steinparzer F., Hiemesch D., Kranawetter E., Salmansberger M., Stütz W., BMW Motoren GmbH, Steyr: The Technical Concept of the New BMW 6-Cylinder 2nd Generation Modular Diesel Engines, Reports 41st

International Vienna Motor Symposium, 22–24 April 2020,

International Vienna Motor Symposium, 22–24 April 2020,

[11] Dr. T. Schell, Mercedes-Benz AG, Stuttgart: M254 – the Future of the 4- Cylinder Gasoline Engine, Reports 41st

[12] Helbing C., Köhne M., Kassel T., Wietholt B., Krause A., Lohre L., Gerhardt N., Eiglmeier C., Volkswagen AG, Wolfsburg:

Volkswagen's TDI-Engines for Euro 6d – Clean Efficiency for Modern Mobility, Reports 41st International Vienna Motor

[13] Schwieberdingen; Univ.-Prof. Dr. H. Eichlseder, Dr. P. Grabner, Dr. K.

Symposium, 22–24 April 2020,

Schaffer, Graz University of Technology: H2 ICE for Future Passenger Cars and Light Commercial Vehicles, Reports 41st International Vienna Motor Symposium, 22–24

[14] Korn T., KEYOU GmbH,

Unterschleißheim: The Most Efficient Way for CO2 Reduction: the New Generation of Hydrogen Internal Combustion Engines, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

[15] Lozanovski A., Geß A., University of Stuttgart; Dipl.-Ing. O. Dingel,

April 2020,

[2] Wisłocki K., Wolański P., Ecker H., Lundqvist U., Pearson R.J., Hartland J.,

Wyszyński M., Powertrain development

discussions at the second International PTNSS Congress, Combustion engines

[3] Lenz H.P., 30 International Vienna Motor Symposium. 7–8 May 2009 – Report on the occasion of the International Congress PTNSS on Combustion Engines 2009 in Opole, Combustion Engines 2/

[4] Walsch M.P., Global trends in motor vehicle pollution control: a 2011 update – part 3, Combustion Engines 4/2011

[5] Mayersohn N. The Internal Combustion Engine Is Not Dead Yet, The New York Times Magazine, 17th

[6] Geringer B., Lenz H.P.,41st International Vienna Motor

[8] Kitadani H., Kaneda R.,

**21**

Symposium, 22–24 April 2020, Reports

[7] Ruhland H., Wirth M., Friedfeld R., Linsel J., Weber C., Krämer F., Ford Werke GmbH, Cologne; Abkenar F., Ford Motor Company, Dearborn, USA: EcoBoost 500: Taking Award Winning Technology to the Next Level, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

Mizoguchi S., Shinohara Y., Takeuchi J., Toyota Motor Corporation, Toyota, Japan: The New 1.5 Liter Gasoline Engine from the TNGA Series, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

Biernat K., Czerwinski J.,

2/2007 (129), 38–53

2009 (137), 150–154.

(167), 98–103.

August 2017

from the perspective of panel

### **Nomenclature**


### **Author details**

Zbigniew J. Sroka Wroclaw University of Science and Technology, Faculty of Mechanical Engineering, Department of Automotive Engineering, Wroclaw, Poland

\*Address all correspondence to: zbigniew.sroka@pwr.edu.pl

© 2021 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

*Work Cycle of Internal Combustion Engine Due to Rightsizing DOI: http://dx.doi.org/10.5772/intechopen.97144*

#### **References**

**Nomenclature**

A coefficient of change of piston stroke B coefficient of change of cylinder diameter

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

Dd cylinder diameter in the downsized engine

Sd stroke of the piston in the downsized engine

BMEP = pe brake mean effective pressure BSFC = ge brake specific fuel consumption D cylinder diameter - input state

Ge fuel consumption per hour Lp moles index for ambient air MR universal gas constant

S stroke of the piston - input state

Wd downsizing/rightsizing index

ε geometric compression ratio ε<sup>s</sup> effective compression ratio

δ degree of another expansion process

η<sup>t</sup> theoretical efficiency of the work cycle

λ<sup>p</sup> degree of pressure increase during isochoric heat transfer ρ' degree of pre-compression when heat is drained at constant

Wroclaw University of Science and Technology, Faculty of Mechanical Engineering, Department of Automotive Engineering, Wroclaw, Poland

© 2021 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium,

\*Address all correspondence to: zbigniew.sroka@pwr.edu.pl

provided the original work is properly cited.

ρ<sup>p</sup> degree of expansion during isobaric heat transfer ρ<sup>T</sup> degree of expansion during isothermal heat transfer

To ambient temperature Vss engine swept volume

Wu fuel calorific value

κ isentropic exponent

**Author details**

Zbigniew J. Sroka

**20**

pressure

τ stroke index (number of strokes)

Ne useful power po ambient pressure rpm = n engine revolution [1] Warnecke W., Lueke W., Clarke L., Louis J., Kempsel S., Fuels of the Future. Proceedings of 27th International Vienna Motor Symposium, Vienna 2006.

[2] Wisłocki K., Wolański P., Ecker H., Lundqvist U., Pearson R.J., Hartland J., Biernat K., Czerwinski J., Wyszyński M., Powertrain development from the perspective of panel discussions at the second International PTNSS Congress, Combustion engines 2/2007 (129), 38–53

[3] Lenz H.P., 30 International Vienna Motor Symposium. 7–8 May 2009 – Report on the occasion of the International Congress PTNSS on Combustion Engines 2009 in Opole, Combustion Engines 2/ 2009 (137), 150–154.

[4] Walsch M.P., Global trends in motor vehicle pollution control: a 2011 update – part 3, Combustion Engines 4/2011 (167), 98–103.

[5] Mayersohn N. The Internal Combustion Engine Is Not Dead Yet, The New York Times Magazine, 17th August 2017

[6] Geringer B., Lenz H.P.,41st International Vienna Motor Symposium, 22–24 April 2020, Reports

[7] Ruhland H., Wirth M., Friedfeld R., Linsel J., Weber C., Krämer F., Ford Werke GmbH, Cologne; Abkenar F., Ford Motor Company, Dearborn, USA: EcoBoost 500: Taking Award Winning Technology to the Next Level, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

[8] Kitadani H., Kaneda R., Mizoguchi S., Shinohara Y., Takeuchi J., Toyota Motor Corporation, Toyota, Japan: The New 1.5 Liter Gasoline Engine from the TNGA Series, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

[9] Song D., Hycet e-Chuang, Great Wall Motor, Hebei, China; W. Happenhofer, Great Wall Motor, Hebei, China: 1.5T High Thermal Efficiency Modular Engine Platform, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

[10] Steinparzer F., Hiemesch D., Kranawetter E., Salmansberger M., Stütz W., BMW Motoren GmbH, Steyr: The Technical Concept of the New BMW 6-Cylinder 2nd Generation Modular Diesel Engines, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

[11] Dr. T. Schell, Mercedes-Benz AG, Stuttgart: M254 – the Future of the 4- Cylinder Gasoline Engine, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

[12] Helbing C., Köhne M., Kassel T., Wietholt B., Krause A., Lohre L., Gerhardt N., Eiglmeier C., Volkswagen AG, Wolfsburg: Volkswagen's TDI-Engines for Euro 6d – Clean Efficiency for Modern Mobility, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

[13] Schwieberdingen; Univ.-Prof. Dr. H. Eichlseder, Dr. P. Grabner, Dr. K. Schaffer, Graz University of Technology: H2 ICE for Future Passenger Cars and Light Commercial Vehicles, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

[14] Korn T., KEYOU GmbH, Unterschleißheim: The Most Efficient Way for CO2 Reduction: the New Generation of Hydrogen Internal Combustion Engines, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

[15] Lozanovski A., Geß A., University of Stuttgart; Dipl.-Ing. O. Dingel,

Dipl.-Ing. (FH) T. Semper, IAV GmbH, Chemnitz: Technical Evaluation and Life Cycle Assessment of Potential Long Haul Heavy Duty Vehicles for the Year 2050, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

[16] Pischinger S. - RWTH Aachen University; van der Put D., Heuser P. - FEV Group GmbH, Aachen; Lindemann B., Müther M., Schönen M. - FEV Europe GmbH, Aachen: Efficient Commercial Powertrains – How to Achieve a 30% GHG Reduction in 2030, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

[17] Hartung S., Member of the Board of Management, Chairman Business Sector Mobility Solutions, Robert Bosch GmbH, Stuttgart: Powertrains of the Future – Sustainable, Safe, Exciting, Reports 41st International Vienna Motor Symposium, 22–24 April 2020,

[18] Heywood J.B., Internal Combustion Engine Fundamentals, McGraw HiU International Editions 1989.

[19] Sroka Z.J., Wybrane zagadnienia teorii tłokowych silni-ków spalinowych w aspekcie zmian objetości skokowej, Oficyna Wydawnicza Politechniki Wrocławskiej, 2013

[20] Fraser A.D.J., How Low can we go? Challenges and opportunities of Engine Downsizing to reduce CO2 Emissions, Seminar Proceedings IMechE, London, 9 February 2011, 1–9.

[21] Pielecha I., Cieślik W., Borowski P., et al. Reduc-tion of the number of cylinders in internal combustion engines – contemporary trends in downsizing. Combustion Engines. 2014, ISSN 2300– 9896, 159(4), 12–25.

[22] Pischinger S, Verbrennungskraftmaschinen I, RWTH Aachen, Aachen 2011.

[23] Fraser N., Bassett M., Extreme Engine Downsizing with a single Turbochanger – 100 kW/l and 30 bar BMEP, Seminar Proceedings IMechE, London, February 2011, 31–45.

[32] Lake T., Stokes J., Murphy R., Downsized DI Gasoline Engines for Low CO2. Seminar Proceedings IMechE, Fuel Economy and Engine Downsizing,

*DOI: http://dx.doi.org/10.5772/intechopen.97144*

*Work Cycle of Internal Combustion Engine Due to Rightsizing*

London, 13 May 2004, 49–55.

[33] Wijetunge R., Criddle M., Dixon J., Morris G., Retaining Driveability in Aggressively Downsized Diesel Engines. Seminar Proceedings IMechE, Fuel Economy and Engine Downsizing, London, 13 May 2004, 41–47.

[34] Mitianiec W., Bac G., Camless hydraulic valve timing system in combustion engines, Combustion Engines 3/2011 (146), 28–37.

[35] www. http://www.ukimediaevents.

[36] Ambrozik A., Wybrane zagadnienia procesów cieplnych w tłokowych silnikach spalinowych, Wydawnictwo

[37] Ambrozik A., Analiza cykli pracy czterosuwowych silników spalinowych,

[38] Blair G.P., Design and Simulation of

com/engineoftheyear/

Kielce 2003.

(3), 2019

**23**

Politechniki Świętokrzyskiej,

Wydawnictwo Politechniki Świętokrzyskiej, Kielce 2010.

Four-Stroke Engine, Society of

[39] Sroka Z.J., Dworaczyński M., Assessment of thermodynamic cycle of internal combustion engine in terms of rightsizing, Combustion Engines, 178

[40] Hordecki J. Volkswagen Golf 1.5TSI – odwórcenie trendu, https:// www/auto-swiat.pl (13 Feb, 2017)

Automotive Engineers, Warrendale 1999.

[24] Jenteges M., van der Weem D., et al. Optimized Activation of a Downsizing Concept with Electrical Boost, MTZ 04/ 2006, Vol. 67.

[25] King J., Application of Synergistic Technologies to Achive High Levels of Gasoline Engine Downsizing, Seminar Proceedings IMechE, London, 9 February 2011, 59–72.

[26] Fisher C.H., Carburation, Vol. III, Spark-Iquition Engines Fuel Injection Systems, Chapman & Hall, London 1966.

[27] Lejda K., Woś P., Fuel Injection in Automotive Engineering – simulation of combustion process in direct injection diesel engine based on fuel injection characteristics, InTech., 2012.

[28] Lejda K., Injection systems of highspeed diesel engines and development trends, Combustion Engines 4/2005 (123), 19–30.

[29] King J., Application of Synergistic Technologies to Achive High Levels of Gasoline Engine Downsizing, Seminar Proceedings IMechE, London, 9 February 2011, 59–72.

[30] Wendeker M., Adaptacyjne sterowanie wtryskiem benzyny w silniku, Państwowe Wydawnictwa Naukowe, Warszawa 2000.

[31] Kammeyer J., Natkaniec C., Seume J.R., Influence of tip-qap losses on the stage efficiency of downsizing turbochanger turbines. Proceedings of 9th International Conference on Turbochangers and Turbocharging (IMechE) 10.1243/ 17547164C0012010023, London, 19–20 May 2010, 293–306

*Work Cycle of Internal Combustion Engine Due to Rightsizing DOI: http://dx.doi.org/10.5772/intechopen.97144*

[32] Lake T., Stokes J., Murphy R., Downsized DI Gasoline Engines for Low CO2. Seminar Proceedings IMechE, Fuel Economy and Engine Downsizing, London, 13 May 2004, 49–55.

Dipl.-Ing. (FH) T. Semper, IAV GmbH, Chemnitz: Technical Evaluation and Life Cycle Assessment of Potential Long Haul Heavy Duty Vehicles for the Year 2050, Reports 41st International Vienna Motor Symposium, 22–24

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

[23] Fraser N., Bassett M., Extreme Engine Downsizing with a single Turbochanger – 100 kW/l and 30 bar BMEP, Seminar Proceedings IMechE, London, February 2011, 31–45.

[24] Jenteges M., van der Weem D., et al. Optimized Activation of a Downsizing Concept with Electrical Boost, MTZ 04/

[25] King J., Application of Synergistic Technologies to Achive High Levels of Gasoline Engine Downsizing, Seminar Proceedings IMechE, London, 9

[26] Fisher C.H., Carburation, Vol. III, Spark-Iquition Engines Fuel Injection

[27] Lejda K., Woś P., Fuel Injection in Automotive Engineering – simulation of combustion process in direct injection diesel engine based on fuel injection characteristics, InTech., 2012.

[28] Lejda K., Injection systems of highspeed diesel engines and development trends, Combustion Engines 4/2005

[29] King J., Application of Synergistic Technologies to Achive High Levels of Gasoline Engine Downsizing, Seminar Proceedings IMechE, London, 9

2006, Vol. 67.

London 1966.

(123), 19–30.

February 2011, 59–72.

(IMechE) 10.1243/

May 2010, 293–306

[30] Wendeker M., Adaptacyjne sterowanie wtryskiem benzyny w silniku, Państwowe Wydawnictwa Naukowe, Warszawa 2000.

[31] Kammeyer J., Natkaniec C., Seume J.R., Influence of tip-qap losses on the stage efficiency of downsizing turbochanger turbines. Proceedings of 9th International Conference on Turbochangers and Turbocharging

17547164C0012010023, London, 19–20

February 2011, 59–72.

Systems, Chapman & Hall,

[16] Pischinger S. - RWTH Aachen University; van der Put D., Heuser P. -

Lindemann B., Müther M., Schönen M. - FEV Europe GmbH, Aachen: Efficient Commercial Powertrains – How to Achieve a 30% GHG Reduction in 2030, Reports 41st International Vienna Motor

[17] Hartung S., Member of the Board of Management, Chairman Business Sector Mobility Solutions, Robert Bosch GmbH, Stuttgart: Powertrains of the Future – Sustainable, Safe, Exciting, Reports 41st International Vienna Motor

[18] Heywood J.B., Internal Combustion Engine Fundamentals, McGraw HiU

[19] Sroka Z.J., Wybrane zagadnienia teorii tłokowych silni-ków spalinowych w aspekcie zmian objetości skokowej, Oficyna Wydawnicza Politechniki

[20] Fraser A.D.J., How Low can we go? Challenges and opportunities of Engine Downsizing to reduce CO2 Emissions, Seminar Proceedings IMechE, London,

[21] Pielecha I., Cieślik W., Borowski P., et al. Reduc-tion of the number of cylinders in internal combustion engines – contemporary trends in downsizing. Combustion Engines. 2014, ISSN 2300–

Verbrennungskraftmaschinen I, RWTH

FEV Group GmbH, Aachen;

Symposium, 22–24 April 2020,

Symposium, 22–24 April 2020,

International Editions 1989.

Wrocławskiej, 2013

9 February 2011, 1–9.

9896, 159(4), 12–25.

Aachen, Aachen 2011.

[22] Pischinger S,

**22**

April 2020,

[33] Wijetunge R., Criddle M., Dixon J., Morris G., Retaining Driveability in Aggressively Downsized Diesel Engines. Seminar Proceedings IMechE, Fuel Economy and Engine Downsizing, London, 13 May 2004, 41–47.

[34] Mitianiec W., Bac G., Camless hydraulic valve timing system in combustion engines, Combustion Engines 3/2011 (146), 28–37.

[35] www. http://www.ukimediaevents. com/engineoftheyear/

[36] Ambrozik A., Wybrane zagadnienia procesów cieplnych w tłokowych silnikach spalinowych, Wydawnictwo Politechniki Świętokrzyskiej, Kielce 2003.

[37] Ambrozik A., Analiza cykli pracy czterosuwowych silników spalinowych, Wydawnictwo Politechniki Świętokrzyskiej, Kielce 2010.

[38] Blair G.P., Design and Simulation of Four-Stroke Engine, Society of Automotive Engineers, Warrendale 1999.

[39] Sroka Z.J., Dworaczyński M., Assessment of thermodynamic cycle of internal combustion engine in terms of rightsizing, Combustion Engines, 178 (3), 2019

[40] Hordecki J. Volkswagen Golf 1.5TSI – odwórcenie trendu, https:// www/auto-swiat.pl (13 Feb, 2017)

**25**

**Chapter 2**

*Hongliang Luo*

**Abstract**

**1. Introduction**

increasing of fuel film on the wall.

Experimental Investigations on

Fuel Spray and Impingement for

Gasoline Direct Injection Engines

Spray-wall impingement is a widespread phenomenon applied in many fields, including spray-wall cooling system, spray coating process and fuel spray and atomization in internal combustion engines. In direct-injection spark ignition (DISI), it is difficult to avoid the fuel film on the piston head and cylinder surfaces. The wet wall caused by impingement affects the air-fuel mixture formation process, which finally influence the subsequent combustion efficiency and performance. Therefore, the fuel spray and impingement under gasoline engine-like conditions were characterized. Mie scattering technique was applied to visualize the spray evolution and impingement processes in a high-pressure and high-temperature constant chamber. Meanwhile, the adhered fuel film on the wall was measured by refractive index matching (RIM) under non-evaporation and evaporation conditions considering the effects of different injection pressures, ambient pressures and ambient temperatures. Additionally, the fuel film formation and evaporation

evolution models were proposed with the help of these mechanisms.

**Keywords:** fuel spray, impingement, fuel film, non-evaporation and evaporation

Generally, compared to port fuel injection (PFI) engine, direct-injection spark ignition (DISI) engines as a greatly potential alternative become more and more widely used for their significant advantages [1, 2]. However, owing to the short impingement distance and high injection pressure, spray impingement on the cylinder wall and piston head is quite difficult to avoid. The impingement affects the fuel-air mixture prior to combustion, which is a possible source for unburned hydrocarbon (UHC) and particulate matters (PM) [3]. Tanaka et al. [4] showed the relationship between the deterioration ratio of engine particle number (PN) emissions and fuel film volume, indicating that PN emissions increases with the

As we known, when fuel immediately out of the nozzle hole, the spray breakup occurs. Two processes of fuel breakup can be involved. The first breakup is also called the primary breakup, leading to large droplets and liquid ligaments near the nozzle to form the dense spray [5, 6]. Under the high injection pressure condition, the cavitation and turbulence are generated from the injector holes, which should be the main reason for this mechanism. Then the following breakup process is named as secondary breakup, indicating these existing droplets break up into

#### **Chapter 2**

## Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines

*Hongliang Luo*

#### **Abstract**

Spray-wall impingement is a widespread phenomenon applied in many fields, including spray-wall cooling system, spray coating process and fuel spray and atomization in internal combustion engines. In direct-injection spark ignition (DISI), it is difficult to avoid the fuel film on the piston head and cylinder surfaces. The wet wall caused by impingement affects the air-fuel mixture formation process, which finally influence the subsequent combustion efficiency and performance. Therefore, the fuel spray and impingement under gasoline engine-like conditions were characterized. Mie scattering technique was applied to visualize the spray evolution and impingement processes in a high-pressure and high-temperature constant chamber. Meanwhile, the adhered fuel film on the wall was measured by refractive index matching (RIM) under non-evaporation and evaporation conditions considering the effects of different injection pressures, ambient pressures and ambient temperatures. Additionally, the fuel film formation and evaporation evolution models were proposed with the help of these mechanisms.

**Keywords:** fuel spray, impingement, fuel film, non-evaporation and evaporation

#### **1. Introduction**

Generally, compared to port fuel injection (PFI) engine, direct-injection spark ignition (DISI) engines as a greatly potential alternative become more and more widely used for their significant advantages [1, 2]. However, owing to the short impingement distance and high injection pressure, spray impingement on the cylinder wall and piston head is quite difficult to avoid. The impingement affects the fuel-air mixture prior to combustion, which is a possible source for unburned hydrocarbon (UHC) and particulate matters (PM) [3]. Tanaka et al. [4] showed the relationship between the deterioration ratio of engine particle number (PN) emissions and fuel film volume, indicating that PN emissions increases with the increasing of fuel film on the wall.

As we known, when fuel immediately out of the nozzle hole, the spray breakup occurs. Two processes of fuel breakup can be involved. The first breakup is also called the primary breakup, leading to large droplets and liquid ligaments near the nozzle to form the dense spray [5, 6]. Under the high injection pressure condition, the cavitation and turbulence are generated from the injector holes, which should be the main reason for this mechanism. Then the following breakup process is named as secondary breakup, indicating these existing droplets break up into

#### *Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

smaller ones owing to aerodynamic forces caused by the relative velocity between droplets and surrounding ambient gas. During the secondary break-up, more and more liquid droplets are formed and move downstream of the spray. Although considerable researches were done for the spray and atomization mechanisms, the spray and impingement are less discussed, let along the fuel film formation after impingement. By considering the competitions from the pure electric vehicle (EV) and concerns on environmental pollution, it is urgent to understand the interaction of liquid droplet and piston wall thoroughly to improve the spray atomization in engine work process.

The target of this study is to investigate the characteristics of gasoline spray impingement and fuel film formation experimentally. The specific objectives of this study are shown as follows:


In this study, Mie Scattering technique was implemented to obtain the observations of gasoline spray emerging from single hole injector under different conditions. RIM technique was adopted to analyze the formation process of fuel liquid on the wall qualitatively and quantitatively under both non-evaporation and evaporation conditions.

#### **2. Experimental apparatus and measurement methods**

As shown in **Figure 1**, a quartz glass (Sigma Koki, DFSQ1-50CO2) was used as the impingement wall with a diameter at 50 mm and thickness at 2 mm.

**27**

**Table 1.** *Test conditions.*

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines*

The coordinate system is established and defined, and the intersection point *o* of the nozzle center axis and the flat wall is decided as the impingement point. The positive *y* axis is along the lateral direction of the spray after impingement, and the positive *x* axis is pointing out of the figure. The impingement angle was 45 deg. and the impingement distance was 22 mm from the nozzle exit to the impingement point of the flat wall along the spray axis. Moreover, the surface roughness was measured at Ra 7.0 μm by a portable high-performance surface roughness and wavi-

The experiment was performed in a constant high-pressure chamber filled with nitrogen gas. A single hole mini-sac injector was used with length at 0.65 mm and hole diameter at 0.155 mm. The test conditions are listed in **Table 1**. Toluene was selected as a substitute for gasoline. The fuel temperature (before injection) was regulated by a cooling system to maintain it at room temperature. The injection pressure changes among 10, 20 and 30 MPa, resulting in the different duration at 2.9, 2.1, and 1.7 ms for fuel injection, which satisfies the constant injection mass at 4.0 mg by considering the real injection mass in each hole. In order to study the influence of ambient pressure, pressures between 0.15 and 0.74 MPa were tested at evaporation condition (*T*amb = 433 K). Meanwhile, the equivalent non-evaporating conditions (*P*amb = 0.1 MPa, *T*amb = 298 K and *P*amb = 0.5 MPa, *T*amb = 298 K) were

, respectively.

ness measuring instrument (Kosaka Laboratory Ltd., SE300).

determined by maintaining the ambient density at 1.95 and 5.95 kg/m3

*P*amb = 0.74 MPa, *T*sat is 472 K, higher than *T*amb.

**2.1 Mie scattering method**

plane with the camera.

One thing should be pointed out that the saturated temperatures (*T*sat) of toluene under *P*amb = 0.15 MPa is 398 K, and it is cleat to see that *T*sat <*T*amb. While under

In order to observe the spray development process, the Mie scatting experiment was performed, and the specific experimental apparatus are shown in **Figure 2**. A high-speed video camera (Photron FASTCAM SA-Z) was utilized to observe the spray with a frame rate at 20,000 frames per second (fps) and a frame size at 512 × 512 pixels. A xenon lamp (Ushio SX-131 UID501XAMQ ) was set in a direction perpendicular to the camera to illuminate the spray, and it was placed in the same

The spray tip penetration (*S*) and impinging spray height (*H*i) are widely used

**Non-evaporation conditions Evaporation conditions**

to investigate the characteristics of the spray-wall impingement [7, 8]. These values were experimentally obtained from raw images by determining the edge of

Ambient Temperature 298 K 433 K Ambient Pressure 0.1 and 0.5 MPa 0.15 and 0.74 MPa Ambient Density 1.95 and 5.95 kg/m3 1.95 and 5.95 kg/m3

Test Fuel Toluene Fuel Temperature 298 K Injection Mass 4.0 mg Ambient Gas Nitrogen Injection Pressure 10, 20, 30 MPa Injection Duration 2.9, 2.1, 1.7 ms

*DOI: http://dx.doi.org/10.5772/intechopen.95848*

**Figure 1.** *Schematic of injector and flat wall.*

#### *Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines DOI: http://dx.doi.org/10.5772/intechopen.95848*

The coordinate system is established and defined, and the intersection point *o* of the nozzle center axis and the flat wall is decided as the impingement point. The positive *y* axis is along the lateral direction of the spray after impingement, and the positive *x* axis is pointing out of the figure. The impingement angle was 45 deg. and the impingement distance was 22 mm from the nozzle exit to the impingement point of the flat wall along the spray axis. Moreover, the surface roughness was measured at Ra 7.0 μm by a portable high-performance surface roughness and waviness measuring instrument (Kosaka Laboratory Ltd., SE300).

The experiment was performed in a constant high-pressure chamber filled with nitrogen gas. A single hole mini-sac injector was used with length at 0.65 mm and hole diameter at 0.155 mm. The test conditions are listed in **Table 1**. Toluene was selected as a substitute for gasoline. The fuel temperature (before injection) was regulated by a cooling system to maintain it at room temperature. The injection pressure changes among 10, 20 and 30 MPa, resulting in the different duration at 2.9, 2.1, and 1.7 ms for fuel injection, which satisfies the constant injection mass at 4.0 mg by considering the real injection mass in each hole. In order to study the influence of ambient pressure, pressures between 0.15 and 0.74 MPa were tested at evaporation condition (*T*amb = 433 K). Meanwhile, the equivalent non-evaporating conditions (*P*amb = 0.1 MPa, *T*amb = 298 K and *P*amb = 0.5 MPa, *T*amb = 298 K) were determined by maintaining the ambient density at 1.95 and 5.95 kg/m3 , respectively. One thing should be pointed out that the saturated temperatures (*T*sat) of toluene under *P*amb = 0.15 MPa is 398 K, and it is cleat to see that *T*sat <*T*amb. While under *P*amb = 0.74 MPa, *T*sat is 472 K, higher than *T*amb.

#### **2.1 Mie scattering method**

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

engine work process.

tion conditions.

study are shown as follows:

interaction between spray and wall.

**2. Experimental apparatus and measurement methods**

smaller ones owing to aerodynamic forces caused by the relative velocity between droplets and surrounding ambient gas. During the secondary break-up, more and more liquid droplets are formed and move downstream of the spray. Although considerable researches were done for the spray and atomization mechanisms, the spray and impingement are less discussed, let along the fuel film formation after impingement. By considering the competitions from the pure electric vehicle (EV) and concerns on environmental pollution, it is urgent to understand the interaction of liquid droplet and piston wall thoroughly to improve the spray atomization in

The target of this study is to investigate the characteristics of gasoline spray impingement and fuel film formation experimentally. The specific objectives of this

• Analyze the impinging spray evolution characteristics which belongs to the single-hole nozzle under the non-evaporation and evaporation conditions.

• Model the fuel film formation to provide an insightful understanding of the

In this study, Mie Scattering technique was implemented to obtain the observations of gasoline spray emerging from single hole injector under different conditions. RIM technique was adopted to analyze the formation process of fuel liquid on the wall qualitatively and quantitatively under both non-evaporation and evapora-

As shown in **Figure 1**, a quartz glass (Sigma Koki, DFSQ1-50CO2) was used

as the impingement wall with a diameter at 50 mm and thickness at 2 mm.

**26**

**Figure 1.**

*Schematic of injector and flat wall.*

In order to observe the spray development process, the Mie scatting experiment was performed, and the specific experimental apparatus are shown in **Figure 2**. A high-speed video camera (Photron FASTCAM SA-Z) was utilized to observe the spray with a frame rate at 20,000 frames per second (fps) and a frame size at 512 × 512 pixels. A xenon lamp (Ushio SX-131 UID501XAMQ ) was set in a direction perpendicular to the camera to illuminate the spray, and it was placed in the same plane with the camera.

The spray tip penetration (*S*) and impinging spray height (*H*i) are widely used to investigate the characteristics of the spray-wall impingement [7, 8]. These values were experimentally obtained from raw images by determining the edge of


**Table 1.** *Test conditions.*

**Figure 3.** *Experimental extraction of S and Hi.*

the impinging spray using inhouse code created in MATLAB software as shown in **Figure 3**. The dotted line represents the wall surface. *S* and *H*i are defined. *S* is defined as the distance from the nozzle exit to the spray tip. Generally, before impingement, *S* is just the distance from the nozzle exit to the spray tip. However, after wall impingement, *S* is defined as the sum of the distance to wall (*L*w) and the radial distance (*L*r) [9]. *H*i is the maximum distance from the wall surface to the edge of the impinging spray. All the results were calculated five times under each specific set of experimental conditions, and the average values were presented.

#### **2.2 RIM method**

The RIM experimental equipment in the current study is shown in **Figure 4**. Some differences can be found in the optical system with Mie scattering. The xenon lamp was placed at the side window to emit continuous light with an incident angle at 15 deg. Under the flat wall, a reflection mirror was positioned directly. Fuel film images were observed by a high-speed video camera through this mirror. Further, the high-speed video camera was set at a frame rate of 10,000 fps and at a frame size of 512 × 512 pixels.

RIM method is applied to measure the fuel film thickness. The image processing is shown in **Figure 5**. First, the image without fuel film named as "dry image" is acquired. After fuel adhering on the wall, it is subtracted by the "dry image"

**29**

**Figure 5.** *Image processing.*

**3.1 Effect of temperature**

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines*

to obtain the only adhered film image. Using the calibration result, the thickness distribution can be obtained. With the scale (0.106 mm/pixel) got from the observation, the film area is calculated by integrating the available pixels. While, for the mass, the thickness can be added up if larger than 0.1 μm. Then the film mass can be

about RIM method and calibration can be seen in the previous publications [10–13].

To discuss the spray characteristics under different ambient temperatures, the images of spray and fuel film at 1.5 ms and 4.0 ms after start of injection (ASOI)

). Additional details

calculated through the scale and density of toluene (867 kg/m3

**3. Impinging spray under evaporation conditions**

*DOI: http://dx.doi.org/10.5772/intechopen.95848*

*Experimental setup for fuel film measurement.*

**Figure 4.**

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines DOI: http://dx.doi.org/10.5772/intechopen.95848*

**Figure 4.** *Experimental setup for fuel film measurement.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

the impinging spray using inhouse code created in MATLAB software as shown in **Figure 3**. The dotted line represents the wall surface. *S* and *H*i are defined. *S* is defined as the distance from the nozzle exit to the spray tip. Generally, before impingement, *S* is just the distance from the nozzle exit to the spray tip. However, after wall impingement, *S* is defined as the sum of the distance to wall (*L*w) and the radial distance (*L*r) [9]. *H*i is the maximum distance from the wall surface to the edge of the impinging spray. All the results were calculated five times under each specific set of experimental conditions, and the average values were presented.

The RIM experimental equipment in the current study is shown in **Figure 4**. Some differences can be found in the optical system with Mie scattering. The xenon lamp was placed at the side window to emit continuous light with an incident angle at 15 deg. Under the flat wall, a reflection mirror was positioned directly. Fuel film images were observed by a high-speed video camera through this mirror. Further, the high-speed video camera was set at a frame rate of 10,000 fps and at a frame

RIM method is applied to measure the fuel film thickness. The image processing is shown in **Figure 5**. First, the image without fuel film named as "dry image" is acquired. After fuel adhering on the wall, it is subtracted by the "dry image"

**28**

**2.2 RIM method**

**Figure 2.**

**Figure 3.**

*Experimental extraction of S and Hi.*

*Experimental apparatus.*

size of 512 × 512 pixels.

to obtain the only adhered film image. Using the calibration result, the thickness distribution can be obtained. With the scale (0.106 mm/pixel) got from the observation, the film area is calculated by integrating the available pixels. While, for the mass, the thickness can be added up if larger than 0.1 μm. Then the film mass can be calculated through the scale and density of toluene (867 kg/m3 ). Additional details about RIM method and calibration can be seen in the previous publications [10–13].

#### **3. Impinging spray under evaporation conditions**

#### **3.1 Effect of temperature**

To discuss the spray characteristics under different ambient temperatures, the images of spray and fuel film at 1.5 ms and 4.0 ms after start of injection (ASOI)

under *T*amb = 298 K are presented in **Figure 6(a)**. The sprays are displayed as binary images acquired from the Mie scattering experiment, and the fuel films are shown in pseudo-color images obtained from the RIM experiment. During injection, the fuel film is incomplete and invalid owing to the mie scattering effect caused by the illumination of the droplets above the wall. While, the fuel film becomes complete and thus available after injection. Therefore, all the results related to the adhesion are shown and discussed after fuel injection. Besides, it is clear to see that the impinging spray height at 4.0 ms ASOI is a little larger than that at 1.5 ms ASOI. And more droplets accumulated at the downstream should be one possible reason for it. Furthermore, the fuel film at impingement region (called Region I) is thicker than that at the periphery region (called Region II) and the thickness decreases along *y* direction. This is due to different mechanisms in the formation of fuel film, which will be explained detailly in the following section.

The images of spray and fuel film at 1.5 ms and 4.0 ms ASOI under *T*amb = 433 K can be seen in **Figure 6(b)**. Under condition of high ambient temperature, the impinging spray height is shorter than that under *T*amb = 298 K, owing to the evaporation of fuel droplets. This phenomenon is more obvious at 4.0 ms AOSI. Moreover, the spray area is smaller in contrast to that observed under *T*amb = 298 K, which can be also explained by the evaporation of droplets during spray. By comparing **Figure 6(a)** and **(b)**, it is evident that evaporation is more significant at the periphery than that observed on Region I.

**Figure 6.** *Spray impingement and fuel film during and after injection. (a) Tamb = 298 K. (b) Tamb = 433 K.*

**31**

**Figure 7.**

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines*

For clarifying the effect of ambient temperature on fuel film thickness distribution along different lines, which were clearly defined to express them concisely. As illustrated in **Figure 7**, the coordinate system is created by setting the impingement point as the origin *o*. As fuel film is almost symmetric along the *y* axis, thickness distribution can be only discussed at different *y* lines at *y* = −2.5, 2.5, 7.5, and 12.5 mm, and the thickness at these lines was described as *H*−2.5, *H*2.5, *H*7.5, and *H*12.5, respectively. The results of *H*−2.5, *H*2.5, *H*7.5, and *H*12.5 at 5 and 50 ms ASOI under *T*amb = 298 K and 433 K are depicted in **Figure 8**, respectively. The horizontal axis is

from −15 to 15 mm, and the vertical axis denotes fuel film thickness.

The results at 5 ms ASOI are shown in **Figure 8(a)**. The thickness under *T*amb = 298 K is larger than that under *T*amb = 433 K due to non-evaporation. For all cases except of *H*12.5, thickness distribution can be divided into three parts as "initially increases, remains constant, and finally decreases". For the constant value, *H*−2.5 and *H*2.5 under 298 K are similar to those under 433 K, but *H*7.5 under 433 K is slightly smaller than that under 298 K. Then as the constant values under 298 K and 433 K are similar, this region is defined as Region I. While for the other region, whose value is variable is defined as Region II. The boundary to separate Regions I and II can be detected through the constant value in thickness using our in-house code made in the MATLAB, which is marked in **Figure 7(a)**. And it is depicted that the Region I under 298 K is almost the same as that under 433 K. A similar observation can be seen in **Figure 7(b)**, although difference can be seen in *H*−2.5, as *H*−2.5 under 433 K evaporates for a quite long time at 50 ms ASOI, leading to a sharp reduction. Regions I and II are presented in **Figure 7(b)** by using the same criterion.

It is worth noting that Region I at 5 ms and 50 ms ASOI are similar as well.

on the wall directly, leading to some fuel sticking on it to form the film.

ited on the wall to form the film because of coalescence and air force.

*Fuel film under different ambient temperatures. (a) 5 ms ASOI (b) 50 ms ASOI.*

applying the divided Regions I and II as mentioned above.

Overall, the mechanisms of fuel film formation are illustrated in **Figure 9** by

Region I also can be named as primary impingement region. The spray impinges

Region II also can be named as secondary impingement region. After fuel spray impingement, most of the fuel splashes off it. The splashing droplets are re-depos-

During injection: When fuel is ejected from the nozzle, liquid fuel breaks up into

ligaments and large droplets. Through the spray propagation, the ligaments and large droplets break up into small ones owing to the interaction between the liquid fuel and ambient gas. In the case of *T*amb = 298 K, after the primary impacting, some liquid fuel deposits on the wall to form Region I, and other fuel splashes off the wall. However, the splashed droplets may collide and coalesce with others, causing a change in velocity. Finally, these droplets may redeposit on the wall to form Region II. Furthermore, under high temperature conditions, less fuel can be seen on Region II,

*DOI: http://dx.doi.org/10.5772/intechopen.95848*

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines DOI: http://dx.doi.org/10.5772/intechopen.95848*

For clarifying the effect of ambient temperature on fuel film thickness distribution along different lines, which were clearly defined to express them concisely. As illustrated in **Figure 7**, the coordinate system is created by setting the impingement point as the origin *o*. As fuel film is almost symmetric along the *y* axis, thickness distribution can be only discussed at different *y* lines at *y* = −2.5, 2.5, 7.5, and 12.5 mm, and the thickness at these lines was described as *H*−2.5, *H*2.5, *H*7.5, and *H*12.5, respectively. The results of *H*−2.5, *H*2.5, *H*7.5, and *H*12.5 at 5 and 50 ms ASOI under *T*amb = 298 K and 433 K are depicted in **Figure 8**, respectively. The horizontal axis is from −15 to 15 mm, and the vertical axis denotes fuel film thickness.

The results at 5 ms ASOI are shown in **Figure 8(a)**. The thickness under *T*amb = 298 K is larger than that under *T*amb = 433 K due to non-evaporation. For all cases except of *H*12.5, thickness distribution can be divided into three parts as "initially increases, remains constant, and finally decreases". For the constant value, *H*−2.5 and *H*2.5 under 298 K are similar to those under 433 K, but *H*7.5 under 433 K is slightly smaller than that under 298 K. Then as the constant values under 298 K and 433 K are similar, this region is defined as Region I. While for the other region, whose value is variable is defined as Region II. The boundary to separate Regions I and II can be detected through the constant value in thickness using our in-house code made in the MATLAB, which is marked in **Figure 7(a)**. And it is depicted that the Region I under 298 K is almost the same as that under 433 K. A similar observation can be seen in **Figure 7(b)**, although difference can be seen in *H*−2.5, as *H*−2.5 under 433 K evaporates for a quite long time at 50 ms ASOI, leading to a sharp reduction. Regions I and II are presented in **Figure 7(b)** by using the same criterion. It is worth noting that Region I at 5 ms and 50 ms ASOI are similar as well.

Overall, the mechanisms of fuel film formation are illustrated in **Figure 9** by applying the divided Regions I and II as mentioned above.

Region I also can be named as primary impingement region. The spray impinges on the wall directly, leading to some fuel sticking on it to form the film.

Region II also can be named as secondary impingement region. After fuel spray impingement, most of the fuel splashes off it. The splashing droplets are re-deposited on the wall to form the film because of coalescence and air force.

During injection: When fuel is ejected from the nozzle, liquid fuel breaks up into ligaments and large droplets. Through the spray propagation, the ligaments and large droplets break up into small ones owing to the interaction between the liquid fuel and ambient gas. In the case of *T*amb = 298 K, after the primary impacting, some liquid fuel deposits on the wall to form Region I, and other fuel splashes off the wall. However, the splashed droplets may collide and coalesce with others, causing a change in velocity. Finally, these droplets may redeposit on the wall to form Region II. Furthermore, under high temperature conditions, less fuel can be seen on Region II,

**Figure 7.** *Fuel film under different ambient temperatures. (a) 5 ms ASOI (b) 50 ms ASOI.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

detailly in the following section.

periphery than that observed on Region I.

under *T*amb = 298 K are presented in **Figure 6(a)**. The sprays are displayed as binary images acquired from the Mie scattering experiment, and the fuel films are shown in pseudo-color images obtained from the RIM experiment. During injection, the fuel film is incomplete and invalid owing to the mie scattering effect caused by the illumination of the droplets above the wall. While, the fuel film becomes complete and thus available after injection. Therefore, all the results related to the adhesion are shown and discussed after fuel injection. Besides, it is clear to see that the impinging spray height at 4.0 ms ASOI is a little larger than that at 1.5 ms ASOI. And more droplets accumulated at the downstream should be one possible reason for it. Furthermore, the fuel film at impingement region (called Region I) is thicker than that at the periphery region (called Region II) and the thickness decreases along *y* direction. This is due to different mechanisms in the formation of fuel film, which will be explained

The images of spray and fuel film at 1.5 ms and 4.0 ms ASOI under *T*amb = 433 K

can be seen in **Figure 6(b)**. Under condition of high ambient temperature, the impinging spray height is shorter than that under *T*amb = 298 K, owing to the evaporation of fuel droplets. This phenomenon is more obvious at 4.0 ms AOSI. Moreover, the spray area is smaller in contrast to that observed under *T*amb = 298 K, which can be also explained by the evaporation of droplets during spray. By comparing **Figure 6(a)** and **(b)**, it is evident that evaporation is more significant at the

*Spray impingement and fuel film during and after injection. (a) Tamb = 298 K. (b) Tamb = 433 K.*

**30**

**Figure 6.**

**Figure 8.** *Fuel film thickness along different lines. (a) 5 ms ASOI (b) 50 ms ASOI.*

#### **Figure 9.**

*Mechanisms of fuel film formation in different regions. (a) During injection. (b) After injection.*

and the decreases in impinging droplets and fuel film evaporation should be the reasons for it.

After injection: Although no fuel spray can be seen, some tiny droplets are still in the air above the wall. Under *T*amb = 298 K, the entrainment air changes the velocity of the splashing droplets. As a result, these droplets may re-impact on the wall to form Regions I and II. Under *T*amb = 298 K, the same phenomena can be expected. But with considering the evaporation effect, the splashing droplets evaporate so quickly that few droplets can redeposit on the wall to from Region II. The fuel film

**33**

**Figure 10.**

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines*

on Region I is mainly formed by the initial impingement, less effect can be seen on Region I, causing only periphery of film evaporating, which leads to fuel film on

To summarize, the fuel film on Region I is caused by the direct impinging spray, affected less by the high ambient temperature. However, film on Region II is mainly caused by the redeposition of the splashing droplets. Under high temperature condition, the splashing droplets evaporate easily before reattaching the wall. Even some can redeposit to form film on the wall, it evaporates quickly due to the strong heat and mass transfer. As a result, fuel film on Region I is similar under *T*amb = 298 K and 433 K, but varies greatly on Region II, as shown in **Figure 7**.

**Figure 10** shows *S* under different injection pressures varying among 10, 20 and 30 MPa. Results at *T*amb = 298 and 433 K are depicted by solid and open data. The horizontal axis represents time after start of fuel injection, and the vertical axis is *S*. The impingement distance is shown by dotted line. Owing to the fuel evaporation, *S* at *T*amb = 433 K is slightly lower than that at *T*amb = 298 K before impingement. Nevertheless, the difference becomes obvious after impingement, which can be attribute to the impingement facilitating fuel breakup and atomization. Moreover, larger difference in *S* can be found under *P*inj = 10 MPa. And the longer time for spray propagation and evaporation may be one possible explanation for it. Although high injection promotes better atomization, the accelerated fuel spray shortens the spray development time, which might be another reason for explaining larger

*H*i can be seen in **Figure 11**. The dotted line represents the end of injection (EOI) with the injection duration depicted. The impingement timing can be advanced under high injection pressure because of high momentum. At EOI, *H*<sup>i</sup> at *P*inj = 30 MPa is larger than that at 10 MPa under both evaporation and nonevaporation condition. The larger Weber number induced by high injection pressure results in more splashing droplets should be the reason for it. Although high injection pressure promotes better atomization, leading to fast evaporation of fuel, the stronger splashing phenomenon cannot be hindered. Furthermore, after EOI at *T*amb = 433 K, as no fuel supply, *H*i under *P*inj = 30 MPa decreases more sharply

*DOI: http://dx.doi.org/10.5772/intechopen.95848*

**3.2 Effect of injection pressure**

difference existing at low injection pressure.

*Spray tip penetration under different injection pressures.*

Region I almost the same as that under room temperature.

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines DOI: http://dx.doi.org/10.5772/intechopen.95848*

on Region I is mainly formed by the initial impingement, less effect can be seen on Region I, causing only periphery of film evaporating, which leads to fuel film on Region I almost the same as that under room temperature.

To summarize, the fuel film on Region I is caused by the direct impinging spray, affected less by the high ambient temperature. However, film on Region II is mainly caused by the redeposition of the splashing droplets. Under high temperature condition, the splashing droplets evaporate easily before reattaching the wall. Even some can redeposit to form film on the wall, it evaporates quickly due to the strong heat and mass transfer. As a result, fuel film on Region I is similar under *T*amb = 298 K and 433 K, but varies greatly on Region II, as shown in **Figure 7**.

#### **3.2 Effect of injection pressure**

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

*Fuel film thickness along different lines. (a) 5 ms ASOI (b) 50 ms ASOI.*

and the decreases in impinging droplets and fuel film evaporation should be the

*Mechanisms of fuel film formation in different regions. (a) During injection. (b) After injection.*

After injection: Although no fuel spray can be seen, some tiny droplets are still in the air above the wall. Under *T*amb = 298 K, the entrainment air changes the velocity of the splashing droplets. As a result, these droplets may re-impact on the wall to form Regions I and II. Under *T*amb = 298 K, the same phenomena can be expected. But with considering the evaporation effect, the splashing droplets evaporate so quickly that few droplets can redeposit on the wall to from Region II. The fuel film

**32**

reasons for it.

**Figure 9.**

**Figure 8.**

**Figure 10** shows *S* under different injection pressures varying among 10, 20 and 30 MPa. Results at *T*amb = 298 and 433 K are depicted by solid and open data. The horizontal axis represents time after start of fuel injection, and the vertical axis is *S*. The impingement distance is shown by dotted line. Owing to the fuel evaporation, *S* at *T*amb = 433 K is slightly lower than that at *T*amb = 298 K before impingement. Nevertheless, the difference becomes obvious after impingement, which can be attribute to the impingement facilitating fuel breakup and atomization. Moreover, larger difference in *S* can be found under *P*inj = 10 MPa. And the longer time for spray propagation and evaporation may be one possible explanation for it. Although high injection promotes better atomization, the accelerated fuel spray shortens the spray development time, which might be another reason for explaining larger difference existing at low injection pressure.

*H*i can be seen in **Figure 11**. The dotted line represents the end of injection (EOI) with the injection duration depicted. The impingement timing can be advanced under high injection pressure because of high momentum. At EOI, *H*<sup>i</sup> at *P*inj = 30 MPa is larger than that at 10 MPa under both evaporation and nonevaporation condition. The larger Weber number induced by high injection pressure results in more splashing droplets should be the reason for it. Although high injection pressure promotes better atomization, leading to fast evaporation of fuel, the stronger splashing phenomenon cannot be hindered. Furthermore, after EOI at *T*amb = 433 K, as no fuel supply, *H*i under *P*inj = 30 MPa decreases more sharply

**Figure 10.** *Spray tip penetration under different injection pressures.*

**Figure 11.** *Impinging spray height under different injection pressures.*

than that under 10 MPa. And the better atomization and fast evaporation should be responsible for it.

In order to discuss the effect of injection pressure on the fuel film evolution, different timing at 5, 20, and 40 ms ASOI are selected to show at different injection pressures under *T*amb = 298 K and 433 K, as shown **Figure 12**. The cross symbol is used to present the impingement point, and color bar is applied to show the thickness from 0 to 2.0 μm. It is important to see the maximum thickness is only 2 μm, and some reasons can be involved to explain it. Firstly, when liquid droplets impacting on the wall, various behaviors can be concluded such as "stick", "spread" and "splash", which are determined by the Weber number of the incident droplets [14]. Therefore, most droplets after impacting could splash off the wall, left few adhering on the wall. Secondly, the rough surface of the wall also promotes the droplets breakup and splashing of droplets after impingement [15]. Thirdly, the formed fuel film may be destroyed by incoming droplets and carried away by the air flow, leading to the maximum thickness at 2 μm. In addition, Ding et al. [16] and Maligne et al. [17] also obtained the similar observation.

**Figure 12(a)** shows fuel film evolution at different injection pressure from 10 to 30 MPa under *T*amb = 298 K. It is clear to see that the film area increases with higher injection pressure. High injection pressure promotes better atomization, as a result, the splashing droplets may re-impact on the wall with the help of the air entrainment and vortex. Besides, the film on Region I decreases under high injection pressure. It was confirmed by Bai et al. [14, 18] that droplet behaviors change from "stick" to "splash" with lager Weber number. And high injection pressure accelerates velocity thus enlarging Weber number. As a result, more fuel splash off the wall after impingement, leading to less film on Region I. Moreover, the similar conclusion can be drawn from Mie scattering results in **Figures 10** and **11** that both *S* and *H*i increase by elevated injection pressure because of droplets splashing. It should be noted that these splashing droplets may re-deposit on the wall to accumulate the film on Region II. Furthermore, film thickness increases with time on both regions as the secondary breakup droplets re-deposit, again.

**Figure 12(b)** shows fuel film evolution at different injection pressure from 10 to 30 MPa under *T*amb = 433 K. Same as **Figure 12(a)**, film on Region I decrease with an increase in injection pressure as the splashing behaviors. Compared to *T*amb = 298 K, fuel film on Region II decreases significantly owing to the fuel evaporation. Moreover, the thickness only becomes larger on Region I. The competition

**35**

**Figure 12.**

*Fuel film on the wall. (a) Tamb = 298 K. (b) Tamb = 433 K.*

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines*

between droplets depositing and evaporation can be purposed. Maligne et al. [17] observed that film evaporates from thin to thick part due to heat transfer. Under high ambient temperature, the tiny droplets evaporate already before impacting.

*DOI: http://dx.doi.org/10.5772/intechopen.95848*

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines DOI: http://dx.doi.org/10.5772/intechopen.95848*

**Figure 12.** *Fuel film on the wall. (a) Tamb = 298 K. (b) Tamb = 433 K.*

between droplets depositing and evaporation can be purposed. Maligne et al. [17] observed that film evaporates from thin to thick part due to heat transfer. Under high ambient temperature, the tiny droplets evaporate already before impacting.

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

than that under 10 MPa. And the better atomization and fast evaporation should be

In order to discuss the effect of injection pressure on the fuel film evolution, different timing at 5, 20, and 40 ms ASOI are selected to show at different injection pressures under *T*amb = 298 K and 433 K, as shown **Figure 12**. The cross symbol is used to present the impingement point, and color bar is applied to show the thickness from 0 to 2.0 μm. It is important to see the maximum thickness is only 2 μm, and some reasons can be involved to explain it. Firstly, when liquid droplets impacting on the wall, various behaviors can be concluded such as "stick", "spread" and "splash", which are determined by the Weber number of the incident droplets [14].

Therefore, most droplets after impacting could splash off the wall, left

Maligne et al. [17] also obtained the similar observation.

as the secondary breakup droplets re-deposit, again.

few adhering on the wall. Secondly, the rough surface of the wall also promotes the droplets breakup and splashing of droplets after impingement [15]. Thirdly, the formed fuel film may be destroyed by incoming droplets and carried away by the air flow, leading to the maximum thickness at 2 μm. In addition, Ding et al. [16] and

**Figure 12(a)** shows fuel film evolution at different injection pressure from 10 to 30 MPa under *T*amb = 298 K. It is clear to see that the film area increases with higher injection pressure. High injection pressure promotes better atomization, as a result, the splashing droplets may re-impact on the wall with the help of the air entrainment and vortex. Besides, the film on Region I decreases under high injection pressure. It was confirmed by Bai et al. [14, 18] that droplet behaviors change from "stick" to "splash" with lager Weber number. And high injection pressure accelerates velocity thus enlarging Weber number. As a result, more fuel splash off the wall after impingement, leading to less film on Region I. Moreover, the similar conclusion can be drawn from Mie scattering results in **Figures 10** and **11** that both *S* and *H*i increase by elevated injection pressure because of droplets splashing. It should be noted that these splashing droplets may re-deposit on the wall to accumulate the film on Region II. Furthermore, film thickness increases with time on both regions

**Figure 12(b)** shows fuel film evolution at different injection pressure from 10 to 30 MPa under *T*amb = 433 K. Same as **Figure 12(a)**, film on Region I decrease with an increase in injection pressure as the splashing behaviors. Compared to *T*amb = 298 K, fuel film on Region II decreases significantly owing to the fuel evaporation. Moreover, the thickness only becomes larger on Region I. The competition

**34**

responsible for it.

*Impinging spray height under different injection pressures.*

**Figure 11.**

Even the secondary droplets re-deposit on Region II to form film, it evaporates very quickly, thus only droplets accumulated on Region I can be observed owing to less heat transfer.

**Figure 13** shows the fuel film mass with time after start of injection. The left vertical axis represents film mass, and right one represents the ratio of film mass to injection mass. The results under *T*amb = 298 K and 433 K are depicted by solid and open data. As we known, two processed that spray development and fuel film formation can be summarized during spray wall interaction [19]. As the strong splashing phenomenon (see from the results of *S* and *H*i above), the maximum only 13% of fuel adheres on the wall under *T*amb = 298 K. Besides, it provides that film mass increases gradually with time due to the splashing droplets depositing on the wall under *T*amb = 298 K. However, under *T*amb = 433 K, the film mass only at *P*inj = 10 MPa increases slightly with time. The transition changing from "splash" to "stick" tends to be responsible for it. More fuel adheres on Region I after impingement at lower injection pressure, leading to thick film. The splashing droplets re-deposit on Region I, it is difficult to evaporate as thick film existing, thus increasing mass slightly with time. For mass at *P*inj = 20 and 30 MPa, the initial film mass on Region I is less in contrast to *P*inj = 10 MPa. Even if the better breakup droplets re-arrive on the wall to form film, it evaporates immediately, which results in a slight decrease of mass. It is interesting to find that high injection pressure increases film mass under *T*amb = 298 K but decreases film mass under *T*amb = 433 K. And the opposite phenomena can be explained by different effect of injection pressure on fuel film formation under evaporation and non-evaporation conditions. Under *T*amb = 298 K, the better atomized droplets accumulate on the wall easily including the first and second adhering on the wall, thus increasing mass at a high injection pressure. However, under *T*amb = 433 K, the better atomized droplets evaporate easily before impacting on the wall for both incident and re-depositing droplets, leading to few film lefts. Besides, high-speed airflow also facilitates the film evaporation on the wall to decrease it.

**Figure 14** describes the film area, and solid and open data are presented to show the results under *T*amb = 298 K and 433 K, respectively. Similar to mass, film area increases with high injection pressure under *T*amb = 298 K, but it decreases under *T*amb = 433 K, which can be attribute to the different effects of injection pressure under ambient temperature conditions. More importantly, as the film evaporates from thin to thick, film area decreases with time, which is different to mass under *P*inj = 10 MPa.

**37**

be detected.

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines*

The film mass distribution at the thickness was defined for deeply study. As illustrated in **Figure 15**, fuel film thickness is presented as the horizontal axis, and film mass is shown by vertical axis. The film mass is calculated by each pixel as

> ( ) ( ) <sup>∞</sup> = =∑*i* 0

where the sum of fuel film mass in the thickness fraction between h ± 0.5 Δ *h*

**Figure 15** shows the film mass along thickness under different conditions at 40 ms ASOI. The peak value decreases dramatically, and curve becomes short owing to the evaporation effect when comparing the results of *T*amb = 298 K and 433 K. It shows that mass under *T*amb = 433 K at the thickness from 0.1–0.8 μm decreases to almost 0 in contrast to that under *T*amb = 298 K, which can be the evidence that high temperature contributes more effort on the thin film evaporation than thick one, thus improving the film uniformity as the narrow shape under *T*amb = 433 K, which is consistent of the observation in **Figure 12**. Besides, the role of injection pressure can be explained more by checking the smallest peak value at

*P*inj = 10 MPa under *T*amb = 298 K, but the largest ones under *T*amb = 433 K.

*i*

∞ =

should be used, and it must satisfy the normalization conditions:

where the *f*M*(h)* is the probability of *M(h)*.

For better understanding the uniformity of the fuel film, the probability of mass

*M i* ( )

As illustrated in **Figure 16**, the peak value of probability under *T*amb = 433 K is much larger than that of *T*amb = 298 K, indicating more uniform film formed on the wall. No thin film at the periphery should be the reason for it due to evaporation. It is interesting to find that under *T*amb = 433 K, although peak value of mass occurs at *P*inj = 10 MPa, but the probability at *P*inj = 30 MPa becomes the largest one, suggesting high injection promotes film uniformity under evaporation condition. However, under non evaporation condition, less influence of injection pressure can

*f h*

∑ <sup>=</sup> 0

*Mh hi* (1)

1 (2)

*DOI: http://dx.doi.org/10.5772/intechopen.95848*

shown in the equation below:

*Fuel film area under different injection pressures.*

**Figure 14.**

and *h* is defined as *M(h)*, and Δ *h* is 0.05 μm.

**Figure 13.** *Fuel film mass under different injection pressures.*

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines DOI: http://dx.doi.org/10.5772/intechopen.95848*

**Figure 14.** *Fuel film area under different injection pressures.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

heat transfer.

ration on the wall to decrease it.

*P*inj = 10 MPa.

Even the secondary droplets re-deposit on Region II to form film, it evaporates very quickly, thus only droplets accumulated on Region I can be observed owing to less

**Figure 13** shows the fuel film mass with time after start of injection. The left vertical axis represents film mass, and right one represents the ratio of film mass to injection mass. The results under *T*amb = 298 K and 433 K are depicted by solid and open data. As we known, two processed that spray development and fuel film formation can be summarized during spray wall interaction [19]. As the strong splashing phenomenon (see from the results of *S* and *H*i above), the maximum only 13% of fuel adheres on the wall under *T*amb = 298 K. Besides, it provides that film mass increases gradually with time due to the splashing droplets depositing on the wall under *T*amb = 298 K. However, under *T*amb = 433 K, the film mass only at *P*inj = 10 MPa increases slightly with time. The transition changing from "splash" to "stick" tends to be responsible for it. More fuel adheres on Region I after impingement at lower injection pressure, leading to thick film. The splashing droplets re-deposit on Region I, it is difficult to evaporate as thick film existing, thus increasing mass slightly with time. For mass at *P*inj = 20 and 30 MPa, the initial film mass on Region I is less in contrast to *P*inj = 10 MPa. Even if the better breakup droplets re-arrive on the wall to form film, it evaporates immediately, which results in a slight decrease of mass. It is interesting to find that high injection pressure increases film mass under *T*amb = 298 K but decreases film mass under *T*amb = 433 K. And the opposite phenomena can be explained by different effect of injection pressure on fuel film formation under evaporation and non-evaporation conditions. Under *T*amb = 298 K, the better atomized droplets accumulate on the wall easily including the first and second adhering on the wall, thus increasing mass at a high injection pressure. However, under *T*amb = 433 K, the better atomized droplets evaporate easily before impacting on the wall for both incident and re-depositing droplets, leading to few film lefts. Besides, high-speed airflow also facilitates the film evapo-

**Figure 14** describes the film area, and solid and open data are presented to show the results under *T*amb = 298 K and 433 K, respectively. Similar to mass, film area increases with high injection pressure under *T*amb = 298 K, but it decreases under *T*amb = 433 K, which can be attribute to the different effects of injection pressure under ambient temperature conditions. More importantly, as the film evaporates from thin to thick, film area decreases with time, which is different to mass under

**36**

**Figure 13.**

*Fuel film mass under different injection pressures.*

The film mass distribution at the thickness was defined for deeply study. As illustrated in **Figure 15**, fuel film thickness is presented as the horizontal axis, and film mass is shown by vertical axis. The film mass is calculated by each pixel as shown in the equation below:

$$M(h) = \sum\_{i=0}^{\circ} h(i) \tag{1}$$

where the sum of fuel film mass in the thickness fraction between h ± 0.5 Δ *h* and *h* is defined as *M(h)*, and Δ *h* is 0.05 μm.

**Figure 15** shows the film mass along thickness under different conditions at 40 ms ASOI. The peak value decreases dramatically, and curve becomes short owing to the evaporation effect when comparing the results of *T*amb = 298 K and 433 K. It shows that mass under *T*amb = 433 K at the thickness from 0.1–0.8 μm decreases to almost 0 in contrast to that under *T*amb = 298 K, which can be the evidence that high temperature contributes more effort on the thin film evaporation than thick one, thus improving the film uniformity as the narrow shape under *T*amb = 433 K, which is consistent of the observation in **Figure 12**. Besides, the role of injection pressure can be explained more by checking the smallest peak value at *P*inj = 10 MPa under *T*amb = 298 K, but the largest ones under *T*amb = 433 K.

For better understanding the uniformity of the fuel film, the probability of mass should be used, and it must satisfy the normalization conditions:

$$\sum\_{i=0}^{s} f\_{\mathcal{M}}\left(h\_i\right) = \mathbf{1} \tag{2}$$

where the *f*M*(h)* is the probability of *M(h)*.

As illustrated in **Figure 16**, the peak value of probability under *T*amb = 433 K is much larger than that of *T*amb = 298 K, indicating more uniform film formed on the wall. No thin film at the periphery should be the reason for it due to evaporation. It is interesting to find that under *T*amb = 433 K, although peak value of mass occurs at *P*inj = 10 MPa, but the probability at *P*inj = 30 MPa becomes the largest one, suggesting high injection promotes film uniformity under evaporation condition. However, under non evaporation condition, less influence of injection pressure can be detected.

**Figure 15.** *Fuel film mass distribution under different injection pressures.*

**Figure 16.** *Probability of film thickness under different injection pressures.*

#### **3.3 Effect of ambient pressure**

**Figure 17** shows the *S* under different ambient pressures, with results of *T*amb = 298 and 433 K depicted by solid and open symbols. It suggests that the increased ambient pressure decreases *S*, and the raised density of the air results in stronger air-fuel entrainment and less spray momentum should be the reason for that under both non-evaporation and evaporation conditions. *H*i under different conditions are compared in **Figure 18**. At the initial stage from 0.2 to 1.0 ms ASOI, *H*i is larger under low ambient pressure. *H*i generates easily as less resistance from ambient pressure should be the reason for it. While, *H*i becomes larger under high ambient pressure with spray development. The raised ambient gas density from high ambient pressure results in strong air-fuel entrainment. Therefore, more air-fuel mixture can be expected after impingement, thus increasing *H*i under high ambient pressure. In contrast to that, after EOI, the spray propagates along the wall

**39**

Region II.

**Figure 17.**

**Figure 18.**

accumulating *H*i cannot be hindered.

*Impinging spray height under different ambient pressures.*

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines*

quickly with less air drag force under low ambient pressure, leading to no fuel supply. Therefore, *H*i decrease sharply after EOI. Besides, the high ambient temperate promotes this behavior. However, owing to the slow air-fuel movement under high ambient pressure, *H*i can keeps constant and even increases to some extent after EOI, suggesting high ambient pressure enlarges *H*i. One thing should be noted that although the high ambient pressure suppresses *H*i, the behaviors of slow movement

In order to discuss the effect of ambient pressure on the fuel film evolution, different timing at 5, 20, and 40 ms ASOI are selected to show at ambient pressures under *T*amb = 298 K and 433 K, as shown **Figure 19**. The cross symbol is used to present the impingement point, and color bar is applied to show the thickness from 0 to 2.0 μm. It has already illustrated in **Figure 9** that two regions on the fuel film on the wall could be separated with different formation mechanism as Regions I and II. And the ambient pressure has more influences on the film formation on

**Figure 19(a)** illustrates that fuel film on both Regions I and II increases with time elapse, and the re-depositing droplets accumulating on the wall should be

*DOI: http://dx.doi.org/10.5772/intechopen.95848*

*Spray tip penetration under different ambient pressures.*

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines DOI: http://dx.doi.org/10.5772/intechopen.95848*

**Figure 17.** *Spray tip penetration under different ambient pressures.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

*Fuel film mass distribution under different injection pressures.*

**38**

**3.3 Effect of ambient pressure**

*Probability of film thickness under different injection pressures.*

**Figure 15.**

**Figure 16.**

**Figure 17** shows the *S* under different ambient pressures, with results of *T*amb = 298 and 433 K depicted by solid and open symbols. It suggests that the increased ambient pressure decreases *S*, and the raised density of the air results in stronger air-fuel entrainment and less spray momentum should be the reason for that under both non-evaporation and evaporation conditions. *H*i under different conditions are compared in **Figure 18**. At the initial stage from 0.2 to 1.0 ms ASOI, *H*i is larger under low ambient pressure. *H*i generates easily as less resistance from ambient pressure should be the reason for it. While, *H*i becomes larger under high ambient pressure with spray development. The raised ambient gas density from high ambient pressure results in strong air-fuel entrainment. Therefore, more air-fuel mixture can be expected after impingement, thus increasing *H*i under high ambient pressure. In contrast to that, after EOI, the spray propagates along the wall

**Figure 18.** *Impinging spray height under different ambient pressures.*

quickly with less air drag force under low ambient pressure, leading to no fuel supply. Therefore, *H*i decrease sharply after EOI. Besides, the high ambient temperate promotes this behavior. However, owing to the slow air-fuel movement under high ambient pressure, *H*i can keeps constant and even increases to some extent after EOI, suggesting high ambient pressure enlarges *H*i. One thing should be noted that although the high ambient pressure suppresses *H*i, the behaviors of slow movement accumulating *H*i cannot be hindered.

In order to discuss the effect of ambient pressure on the fuel film evolution, different timing at 5, 20, and 40 ms ASOI are selected to show at ambient pressures under *T*amb = 298 K and 433 K, as shown **Figure 19**. The cross symbol is used to present the impingement point, and color bar is applied to show the thickness from 0 to 2.0 μm. It has already illustrated in **Figure 9** that two regions on the fuel film on the wall could be separated with different formation mechanism as Regions I and II. And the ambient pressure has more influences on the film formation on Region II.

**Figure 19(a)** illustrates that fuel film on both Regions I and II increases with time elapse, and the re-depositing droplets accumulating on the wall should be

the reason for it. Moreover, it has to be noted that the film area increases with high ambient pressure. Three main reasons can be summarized. Firstly, elevated ambient pressure promotes better atomization, and these atomized droplets re-arrive on the wall easily by vortex to increase film. Secondly, high ambient pressure decelerates fuel spray before impingement, leading to decreasing Weber number. Therefore, the droplet behaviors may change from "splash" to "spread" or "stick" with small Weber number, causing adhered film increase. Thirdly, under high ambient pressure, the splashing droplets could collide and coalesce with others easily, and finally re-deposit on the wall to increase the fuel film.

**41**

**Figure 20.**

*Fuel film mass under different ambient pressures.*

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines*

**Figure 19(b)** illustrates the fuel film evolution on the wall at *T*amb = 433 K. It is clear that owing to evaporation, fuel film area decreases obviously with time. Furthermore, more fuel locates on Region I other than Region II, and the less heat transfer at thick film should be the reason for it. Under *T*amb = 433 K, some of the splashing droplets evaporate before re-impacting on the wall. And the survived re-depositing droplets on Region II evaporate immediately as strong hear transfer, leading to more fuel film accumulating on Region II at last. While, the survived ones impact on Region I then form film on it successfully. It can be seen that fuel film exist on Region II at 5 ms ASOI under *P*amb = 0.74 MPa, and the large droplets re-depositing and adhering on the wall before evaporation should be the main reason for it. Besides, the fuel film area under *P*amb = 0.74 MPa is larger than that under *P*amb = 0.15 MPa. In addition to the reasons discussed above, high ambient pressure hinders the film evaporation to a certain extent could be another reason. One more interesting thing is that under *P*amb = 0.15 MPa, fuel film evaporates from periphery until 40 ms ASOI, and the "collapse" appearing in the film affects the film uniformity, which will be

**Figure 20** shows the fuel film mass with time after start of injection. The left vertical axis represents film mass, and right one represents the ratio of film mass to injection mass. The results under *T*amb = 298 K and 433 K are depicted by solid and open data. It depicts that film mass increases with time gradually under *T*amb = 298 K. And the re-depositing droplets after splashing should be responsible for it. Noticed that the increased amplitude of film mass under *P*amb = 0.5 MPa is much larger when compared to of *P*amb = 0.1 MPa. On one hand, high ambient pressure enlarges and decelerate droplets, leading to more droplets tend to re-deposit on the wall. On the other hand, the behavior transition from "splash" to "stich" increases film on the wall. But looking at the mass under *T*amb = 433 K, it decreases with time slightly due to evaporation. Even if some re-deposited droplets adhere on the wall, the evaporation effect cannot be ignored. More importantly, mass under *P*amb = 0.74 MPa is larger than that under *P*amb = 0.15 MPa. As explained in the experimental condition, *T*sat (398 K) < *T*amb under *P*amb = 0.15 MPa, indicating a strong film evaporation. While, *T*sat (472 K) is larger than *T*amb under *P*amb = 0.74 MPa, and a slow evaporation can be expected. Recently, Schulz et al. [20] found that the time for film evaporation duration increases at elevated ambient pressure, implying that high

*DOI: http://dx.doi.org/10.5772/intechopen.95848*

discussed more in the following part.

**Figure 19.** *Evolution of fuel film on the wall. (a) Tamb = 298 K. (b) Tamb = 433 K.*

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines DOI: http://dx.doi.org/10.5772/intechopen.95848*

**Figure 19(b)** illustrates the fuel film evolution on the wall at *T*amb = 433 K. It is clear that owing to evaporation, fuel film area decreases obviously with time. Furthermore, more fuel locates on Region I other than Region II, and the less heat transfer at thick film should be the reason for it. Under *T*amb = 433 K, some of the splashing droplets evaporate before re-impacting on the wall. And the survived re-depositing droplets on Region II evaporate immediately as strong hear transfer, leading to more fuel film accumulating on Region II at last. While, the survived ones impact on Region I then form film on it successfully. It can be seen that fuel film exist on Region II at 5 ms ASOI under *P*amb = 0.74 MPa, and the large droplets re-depositing and adhering on the wall before evaporation should be the main reason for it. Besides, the fuel film area under *P*amb = 0.74 MPa is larger than that under *P*amb = 0.15 MPa. In addition to the reasons discussed above, high ambient pressure hinders the film evaporation to a certain extent could be another reason. One more interesting thing is that under *P*amb = 0.15 MPa, fuel film evaporates from periphery until 40 ms ASOI, and the "collapse" appearing in the film affects the film uniformity, which will be discussed more in the following part.

**Figure 20** shows the fuel film mass with time after start of injection. The left vertical axis represents film mass, and right one represents the ratio of film mass to injection mass. The results under *T*amb = 298 K and 433 K are depicted by solid and open data. It depicts that film mass increases with time gradually under *T*amb = 298 K. And the re-depositing droplets after splashing should be responsible for it. Noticed that the increased amplitude of film mass under *P*amb = 0.5 MPa is much larger when compared to of *P*amb = 0.1 MPa. On one hand, high ambient pressure enlarges and decelerate droplets, leading to more droplets tend to re-deposit on the wall. On the other hand, the behavior transition from "splash" to "stich" increases film on the wall. But looking at the mass under *T*amb = 433 K, it decreases with time slightly due to evaporation. Even if some re-deposited droplets adhere on the wall, the evaporation effect cannot be ignored. More importantly, mass under *P*amb = 0.74 MPa is larger than that under *P*amb = 0.15 MPa. As explained in the experimental condition, *T*sat (398 K) < *T*amb under *P*amb = 0.15 MPa, indicating a strong film evaporation. While, *T*sat (472 K) is larger than *T*amb under *P*amb = 0.74 MPa, and a slow evaporation can be expected. Recently, Schulz et al. [20] found that the time for film evaporation duration increases at elevated ambient pressure, implying that high

**Figure 20.** *Fuel film mass under different ambient pressures.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

re-deposit on the wall to increase the fuel film.

the reason for it. Moreover, it has to be noted that the film area increases with high ambient pressure. Three main reasons can be summarized. Firstly, elevated ambient pressure promotes better atomization, and these atomized droplets re-arrive on the wall easily by vortex to increase film. Secondly, high ambient pressure decelerates fuel spray before impingement, leading to decreasing Weber number. Therefore, the droplet behaviors may change from "splash" to "spread" or "stick" with small Weber number, causing adhered film increase. Thirdly, under high ambient pressure, the splashing droplets could collide and coalesce with others easily, and finally

**40**

**Figure 19.**

*Evolution of fuel film on the wall. (a) Tamb = 298 K. (b) Tamb = 433 K.*

ambient pressure has a significant effect on inhibiting evaporation of fuel film. This founding is confirmed by Kim et al. [21] from their single droplet experiment and Tao et al. [22] from the computational fluid dynamics (CFD)results. Furthermore, the results that film mass under *P*amb = 0.74 MPa is larger than that under *P*amb = 0.15 MPa can also demonstrate that high ambient pressure inhibits the fuel evaporation.

**Figure 21** describes the film area, and solid and open data are presented to show the results under *T*amb = 298 K and 433 K, respectively. As the re-depositing droplets impact on the same location, leading to large increase in mass but little increase in area. Therefore, fuel film under *T*amb = 298 K increases with time slightly. Same as film mass, the area increases with high ambient pressure. Under non-evaporation condition, the large size, low velocity, "stick" behavior and easy re-deposition should be attribute to it. Under evaporation condition, in addition to the above factors, the inhibition of film evaporation under high ambient pressure should be another reason for it.

**Figure 22** shows the film mass along thickness under different conditions at 40 ms ASOI. And the results under *T*amb = 298 K and 433 K are shown by solid and

**Figure 21.** *Fuel film area under different ambient pressures.*

**43**

**Figure 23.**

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines*

open symbols, respectively. It reveals that the peak value increases significantly with high ambient pressure. It should be clarified that under *T*amb = 298 K, tow peak values at 0.5 and 1.6 μm can be observed for *P*amb = 0.1 MPa condition, indicating film accumulated on Region II and the central of region I. However, by increasing the ambient pressure to 0.5 MPa, it is interesting to find these two peaks merge into one, indicating more uniform film form under high ambient

**Figure 23** shows the probability of film mass along thickness under different conditions at 40 ms ASOI, and the results under *T*amb = 298 K and 433 K are depicted by solid and open symbols, respectively. The probability increases sharply under evaporation condition, indicating high temperature facilitates uniform film on the wall. And the evaporation occurs from thin to thick should be the main reason. Special attention should be taken that not only ambient temperature, but also ambient pressure can improve the fuel film uniformity. Elevating the ambient pressure at a certain temperature can increase the density of the ambient gas, leading to a strong interaction between air gas and fuel. As a result, fuel spray angle increases, and droplets kinetic energy decreases. Eventually, the droplets attach o more evenly on the wall, thereby forming a uniform film on the wall. Because of this reason, under *T*amb = 298 K, when ambient pressure increased from 0.1 to 0.5 MPa, these two peaks merge into one, indicating large improvement for film uniformity. Similarly, under *T*amb = 433 K, as less film located on Region II, the two peaks disappear. By increasing the ambient pressure from 0.15 to 0.74 MPa, peak value increases, suggesting uniformity improved, although not so obvious as that under

To sum up, **Figure 24** displays the relationship between the saturation vapor pressure curve and the tested conditions. It shows that above the curve toluene is under the liquid phase but transits to vapor phase below the curve. By increasing the ambient pressure from 0.1 to 0.5 MPa under *T*amb = 298 K, the ambient density increases. Then, the stronger entrainment and air resistance can be expected, leading to low velocity and large size of the droplets, and behavior changing from "splash" to "stick", which are the reasons for fuel film increase. In addition, increased the ambient pressure from 0.15 to 0.74 MPa under *T*amb = 433 K, because of the transient from vapor to liquid phase, the decrease in evaporation rate as one

*DOI: http://dx.doi.org/10.5772/intechopen.95848*

pressure.

non-evaporation condition.

more reason increases fuel film on the wall.

*Probability of film thickness under different ambient pressures.*

**Figure 22.** *Fuel film mass distribution under different ambient pressures.*

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines DOI: http://dx.doi.org/10.5772/intechopen.95848*

open symbols, respectively. It reveals that the peak value increases significantly with high ambient pressure. It should be clarified that under *T*amb = 298 K, tow peak values at 0.5 and 1.6 μm can be observed for *P*amb = 0.1 MPa condition, indicating film accumulated on Region II and the central of region I. However, by increasing the ambient pressure to 0.5 MPa, it is interesting to find these two peaks merge into one, indicating more uniform film form under high ambient pressure.

**Figure 23** shows the probability of film mass along thickness under different conditions at 40 ms ASOI, and the results under *T*amb = 298 K and 433 K are depicted by solid and open symbols, respectively. The probability increases sharply under evaporation condition, indicating high temperature facilitates uniform film on the wall. And the evaporation occurs from thin to thick should be the main reason. Special attention should be taken that not only ambient temperature, but also ambient pressure can improve the fuel film uniformity. Elevating the ambient pressure at a certain temperature can increase the density of the ambient gas, leading to a strong interaction between air gas and fuel. As a result, fuel spray angle increases, and droplets kinetic energy decreases. Eventually, the droplets attach o more evenly on the wall, thereby forming a uniform film on the wall. Because of this reason, under *T*amb = 298 K, when ambient pressure increased from 0.1 to 0.5 MPa, these two peaks merge into one, indicating large improvement for film uniformity. Similarly, under *T*amb = 433 K, as less film located on Region II, the two peaks disappear. By increasing the ambient pressure from 0.15 to 0.74 MPa, peak value increases, suggesting uniformity improved, although not so obvious as that under non-evaporation condition.

To sum up, **Figure 24** displays the relationship between the saturation vapor pressure curve and the tested conditions. It shows that above the curve toluene is under the liquid phase but transits to vapor phase below the curve. By increasing the ambient pressure from 0.1 to 0.5 MPa under *T*amb = 298 K, the ambient density increases. Then, the stronger entrainment and air resistance can be expected, leading to low velocity and large size of the droplets, and behavior changing from "splash" to "stick", which are the reasons for fuel film increase. In addition, increased the ambient pressure from 0.15 to 0.74 MPa under *T*amb = 433 K, because of the transient from vapor to liquid phase, the decrease in evaporation rate as one more reason increases fuel film on the wall.

**Figure 23.** *Probability of film thickness under different ambient pressures.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

the fuel evaporation.

ambient pressure has a significant effect on inhibiting evaporation of fuel film. This founding is confirmed by Kim et al. [21] from their single droplet experiment and Tao et al. [22] from the computational fluid dynamics (CFD)results. Furthermore, the results that film mass under *P*amb = 0.74 MPa is larger than that under *P*amb = 0.15 MPa can also demonstrate that high ambient pressure inhibits

**Figure 21** describes the film area, and solid and open data are presented to show the results under *T*amb = 298 K and 433 K, respectively. As the re-depositing droplets impact on the same location, leading to large increase in mass but little increase in area. Therefore, fuel film under *T*amb = 298 K increases with time slightly. Same as film mass, the area increases with high ambient pressure. Under non-evaporation condition, the large size, low velocity, "stick" behavior and easy re-deposition should be attribute to it. Under evaporation condition, in addition to the above factors, the inhibition of film

**Figure 22** shows the film mass along thickness under different conditions at 40 ms ASOI. And the results under *T*amb = 298 K and 433 K are shown by solid and

evaporation under high ambient pressure should be another reason for it.

**42**

**Figure 22.**

**Figure 21.**

*Fuel film area under different ambient pressures.*

*Fuel film mass distribution under different ambient pressures.*

**Figure 24.** *Saturation vapor curve of tested fuel.*

#### **3.4 Fuel film evaporation characteristics**

**Figure 25(a)** shows the fuel film evolution at *P*amb = 0.15 MPa. Different timings at 50, 70, 100, and 150 ms ASOI are selected to depict and the color bar presents thickness from 0 to 2.0 μm with cross symbol being as impingement point (*o*). Owing to the fuel evaporation, the film area decreases with time. Furthermore, the film evaporates from the cross symbol to downstream, making the fuel film "moves downward". The film thickness increases from up to downstream due to the interaction of fuel spray and wall, leading to the fuel film evaporates quickly at the impingement region. As a result, it appears the film moving downward. Besides, the increased injection pressure decreases film area significantly. Three reasons can be concluded. Firstly, the high injection promotes better atomization, resulting in more tiny droplets evaporate before impacting on the wall. Secondly, even if some droplets survive to impact on the wall, with larger Weber number, the trend of "splash" after impingement can be expected. Thirdly, even though fuel finally sticks on the wall, the air flow with high velocity improves the evaporation rate, leading to less film left. Noteworthy is that the film evaporates from the periphery to the central of film firstly as Maligne and Bruneaux [17] reported, then evaporates from the central causing a "hollow" in the film, which will be analyzed in the following part. It is interesting that under the elevated injection pressure, more film can be observed near the cross symbol, especially at 150 ms ASOI. The increased injection pressure enlarges Weber number, leading to more fuel "splash" instead of "spread". Finally, less fuel accumulates at the downstream.

**Figure 25(b)** shows the fuel film evolution at *P*amb = 0.74 MPa. More fuel film can be seen left on the wall when ambient pressure increased to 0.74 MPa. The same as **Figure 25(a)**, it is evident that fuel film area decreases sharply against high injection pressure. In addition, the increased ambient pressure also increases film area. The droplets are decelerated by stronger air force of high ambient pressure. Therefore, more droplets tend to "stick" on the wall, resulting in more film formed. Besides, even some droplets splash off the wall, the splashing droplets tend to collide and then coalesce with others. Finally, these larger droplets redeposit on the wall more easily, thus accumulating more film. Moreover, the evaporation rate decreases under high ambient pressure, leading to more film left. And more discussions about the evaporation rate

**45**

**Figure 25.**

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines*

will be discussed in the next part. Besides, similar to **Figure 25(a)**, more fuel can be seen near the cross symbol with an increase in injection pressure at 500 ms ASOI.

*Evolution of fuel film. (a) Pamb = 0.15 MPa. (b) Pamb = 0.74 MPa.*

being mass and right one being the ratio of film mass to injection mass. Results

**Figure 26** shows fuel film mass under different conditions with left vertical axis

*DOI: http://dx.doi.org/10.5772/intechopen.95848*

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines DOI: http://dx.doi.org/10.5772/intechopen.95848*

**Figure 25.** *Evolution of fuel film. (a) Pamb = 0.15 MPa. (b) Pamb = 0.74 MPa.*

will be discussed in the next part. Besides, similar to **Figure 25(a)**, more fuel can be seen near the cross symbol with an increase in injection pressure at 500 ms ASOI.

**Figure 26** shows fuel film mass under different conditions with left vertical axis being mass and right one being the ratio of film mass to injection mass. Results

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

**3.4 Fuel film evaporation characteristics**

*Saturation vapor curve of tested fuel.*

**Figure 24.**

**Figure 25(a)** shows the fuel film evolution at *P*amb = 0.15 MPa. Different timings at 50, 70, 100, and 150 ms ASOI are selected to depict and the color bar presents thickness from 0 to 2.0 μm with cross symbol being as impingement point (*o*). Owing to the fuel evaporation, the film area decreases with time. Furthermore, the film evaporates from the cross symbol to downstream, making the fuel film "moves downward". The film thickness increases from up to downstream due to the interaction of fuel spray and wall, leading to the fuel film evaporates quickly at the impingement region. As a result, it appears the film moving downward. Besides, the increased injection pressure decreases film area significantly. Three reasons can be concluded. Firstly, the high injection promotes better atomization, resulting in more tiny droplets evaporate before impacting on the wall. Secondly, even if some droplets survive to impact on the wall, with larger Weber number, the trend of "splash" after impingement can be expected. Thirdly, even though fuel finally sticks on the wall, the air flow with high velocity improves the evaporation rate, leading to less film left. Noteworthy is that the film evaporates from the periphery to the central of film firstly as Maligne and Bruneaux [17] reported, then evaporates from the central causing a "hollow" in the film, which will be analyzed in the following part. It is interesting that under the elevated injection pressure, more film can be observed near the cross symbol, especially at 150 ms ASOI. The increased injection pressure enlarges Weber number, leading to more fuel "splash"

instead of "spread". Finally, less fuel accumulates at the downstream.

**Figure 25(b)** shows the fuel film evolution at *P*amb = 0.74 MPa. More fuel film can be seen left on the wall when ambient pressure increased to 0.74 MPa. The same as **Figure 25(a)**, it is evident that fuel film area decreases sharply against high injection pressure. In addition, the increased ambient pressure also increases film area. The droplets are decelerated by stronger air force of high ambient pressure. Therefore, more droplets tend to "stick" on the wall, resulting in more film formed. Besides, even some droplets splash off the wall, the splashing droplets tend to collide and then coalesce with others. Finally, these larger droplets redeposit on the wall more easily, thus accumulating more film. Moreover, the evaporation rate decreases under high ambient pressure, leading to more film left. And more discussions about the evaporation rate

**44**

**Figure 26.** *Evaporation characteristics of fuel film mass.*

under *P*amb = 0.15 and 0.74 MPa are shown by filled and open data. It is evident that the mass increases with an increase in injection pressure due to the better atomization and fast evaporating rate. While, the mass increases with an increase in ambient pressure. More fuel sticking on the wall and low evaporating rate should be responsible for it. It should be noted that the film mass firstly increases and then decreases under only one case of *P*inj = 10 MPa, *P*amb = 0.74 MPa. Although the injection duration is only 2.9 ms, mass increases until 200 ms ASOI. It should be attributed to the lowest velocity of liquid droplets and largest size of them. As described above, under high ambient but low injection pressures condition, the spray is decelerated causing the impacting droplets arrive at the wall relatively late, which is the main reason for mass increasing gradually. Moreover, droplets tend to coalesce into larger more after impingement making it difficult to evaporate. In addition, droplets re-deposit on the wall easily under this case. It is interesting to see that the maximum ratio is approximately 4%, while the minimum one is 1.5%, suggesting most fuel evaporates or splashes off the wall.

**Figure 27** shows fuel film area under different conditions. Results under *P*amb = 0.15 and 0.74 MPa are shown by filled and open data. Similar to mass, the elevated injection pressure enlarges film area decreases but ambient pressure increases it with the same reason as mass. More importantly, different to mass, no obvious increase in area can be seen under *P*inj = 10 MPa, *P*amb = 0.74 MPa, implying that the impacting droplets arrive at the same location, leading to only mass increase. Besides, the lifetime of fuel film can be obtained from **Figures 26** and **27**, which will be discussed more in **Figure 28**.

**Figure 29** provides the details about the "hollow" development in fuel film under *P*inj = 10 MPa, *P*amb = 0.15 MPa. It reveals that film evaporates at periphery from 60 to 120 ms ASOI, causing it "moving downward" as analyzed in **Figure 25**. Generally, evaporation occurs during the film formation, also leading to the film becomes thin with time for the whole region. At around 140 ms, it shows "hollow" occurs. With "hollow" development, film becomes less uniform, leading to quick evaporation at "hollow" until all evaporated.

A conceptual model for fuel film evaporation is proposed here to describe its mechanism and evaporation dynamic, as shown in **Figure 30**. Three processed can be summarized "Evaporate from periphery - Hollow occur - Evaporate from hollow". Because the fuel film at the periphery is relatively thin, strong evaporation can be expected owing to heat transfer in contrast to central part. With the consideration of the flat wall used with roughness, the film is non-uniform initially. Then,

**47**

**Figure 27.**

**Figure 28.** *Fuel film lifetime.*

*Evaporation characteristics of fuel film area.*

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines*

"hollow" appears in the film with evaporation, leading to film "breakup". With the "hollow" development, film evaporation is facilitated, and film becomes less

can be found from the injection pressure on phase changing.

**Figure 28** illustrates the fuel film lifetime under different conditions. Noting that the lifetime is calculated from the start of injection until all film evaporates, as shown in **Figures 26** and **27**. One more thing should be noted that although the lifetime is quite longer than one cycle of the real working condition in gasoline engine, lifetime is an important factor to evaluate film evaporation characteristics, especially for validating the simulation result. The observation shows that injection pressure decreases film lifetime owing to the enhanced evaporation rate and better atomization. Moreover, the decreased lifetime from 10 to 20 MPa is more obvious than that from 20 to 30 MPa, suggesting limitation may exist for the high injection pressure shortening lifetime. But ambient pressure increases the lifetime mainly due to the liquid/vapor transition by elevating ambient pressure from 0.15 to 0.74 MPa. The phase transition can decrease the evaporation rate sharply, leading to longer lifetime of the film. Besides, this may be another reason for the mass increases with high ambient pressure. But no influence

uniform until all evaporates finally.

*DOI: http://dx.doi.org/10.5772/intechopen.95848*

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines DOI: http://dx.doi.org/10.5772/intechopen.95848*

**Figure 27.** *Evaporation characteristics of fuel film area.*

**Figure 28.** *Fuel film lifetime.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

under *P*amb = 0.15 and 0.74 MPa are shown by filled and open data. It is evident that the mass increases with an increase in injection pressure due to the better atomization and fast evaporating rate. While, the mass increases with an increase in ambient pressure. More fuel sticking on the wall and low evaporating rate should be responsible for it. It should be noted that the film mass firstly increases and then decreases under only one case of *P*inj = 10 MPa, *P*amb = 0.74 MPa. Although the injection duration is only 2.9 ms, mass increases until 200 ms ASOI. It should be attributed to the lowest velocity of liquid droplets and largest size of them. As described above, under high ambient but low injection pressures condition, the spray is decelerated causing the impacting droplets arrive at the wall relatively late, which is the main reason for mass increasing gradually. Moreover, droplets tend to coalesce into larger more after impingement making it difficult to evaporate. In addition, droplets re-deposit on the wall easily under this case. It is interesting to see that the maximum ratio is approximately 4%, while the minimum one is 1.5%,

**Figure 27** shows fuel film area under different conditions. Results under *P*amb = 0.15 and 0.74 MPa are shown by filled and open data. Similar to mass, the elevated injection pressure enlarges film area decreases but ambient pressure increases it with the same reason as mass. More importantly, different to mass, no obvious increase in area can be seen under *P*inj = 10 MPa, *P*amb = 0.74 MPa, implying that the impacting droplets arrive at the same location, leading to only mass increase. Besides, the lifetime of fuel film can be obtained from **Figures 26** and **27**,

**Figure 29** provides the details about the "hollow" development in fuel film under *P*inj = 10 MPa, *P*amb = 0.15 MPa. It reveals that film evaporates at periphery from 60 to 120 ms ASOI, causing it "moving downward" as analyzed in **Figure 25**. Generally, evaporation occurs during the film formation, also leading to the film becomes thin with time for the whole region. At around 140 ms, it shows "hollow" occurs. With "hollow" development, film becomes less uniform, leading to quick

A conceptual model for fuel film evaporation is proposed here to describe its mechanism and evaporation dynamic, as shown in **Figure 30**. Three processed can be summarized "Evaporate from periphery - Hollow occur - Evaporate from hollow". Because the fuel film at the periphery is relatively thin, strong evaporation can be expected owing to heat transfer in contrast to central part. With the consideration of the flat wall used with roughness, the film is non-uniform initially. Then,

suggesting most fuel evaporates or splashes off the wall.

which will be discussed more in **Figure 28**.

evaporation at "hollow" until all evaporated.

**46**

**Figure 26.**

*Evaporation characteristics of fuel film mass.*

"hollow" appears in the film with evaporation, leading to film "breakup". With the "hollow" development, film evaporation is facilitated, and film becomes less uniform until all evaporates finally.

**Figure 28** illustrates the fuel film lifetime under different conditions. Noting that the lifetime is calculated from the start of injection until all film evaporates, as shown in **Figures 26** and **27**. One more thing should be noted that although the lifetime is quite longer than one cycle of the real working condition in gasoline engine, lifetime is an important factor to evaluate film evaporation characteristics, especially for validating the simulation result. The observation shows that injection pressure decreases film lifetime owing to the enhanced evaporation rate and better atomization. Moreover, the decreased lifetime from 10 to 20 MPa is more obvious than that from 20 to 30 MPa, suggesting limitation may exist for the high injection pressure shortening lifetime. But ambient pressure increases the lifetime mainly due to the liquid/vapor transition by elevating ambient pressure from 0.15 to 0.74 MPa. The phase transition can decrease the evaporation rate sharply, leading to longer lifetime of the film. Besides, this may be another reason for the mass increases with high ambient pressure. But no influence can be found from the injection pressure on phase changing.

**Figure 29.** *"Hollow" development.*

#### **Figure 30.**

*Schematic conceptual model. (a) Film evaporates from periphery. (b) With evaporation, "hollow"occurs. (c) Film evaporates from "hollow".*

#### **4. Conclusions**

This chapter discussed about the fuel spray and impingement under non-evaporation and evaporation conditions. The impinging spray development and fuel film formation as well as evaporation characteristics were checked and compared, the major conclusions are summarized as follows:

1.Different impingement regions were characterized through the formation of fuel film, known as the primary impingement region (called Region I) and the

**49**

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines*

2.During the injection, a high injection pressure accelerates the fuel spray and favors the atomization of fuel, thus enlarging *S* and *H*i of liquid phase under both *T*amb = 298 K and *T*amb = 433 K. Under *T*amb = 298 K, injection pressure increases the fuel film mass and area. However, under *T*amb = 433 K, injection pressure decreases the them, and the better atomization and easy evaporation

3.The increased ambient pressure increases the density of the gas, resulting in strong air-fuel entertainment and less momentum, thus shortening *S* and increasing *H*i. Moreover, higher ambient pressure increases the fuel film mass

4.The fuel film firstly evaporates from periphery, leading to fuel film "moving downward". Then, "hollow" occurs and develops, resulting in less uniformity of fuel film. Finally, the fuel film evaporates from the "hollow" until all evaporated. The fuel film lifetime is shortened by increasing injection pressure, but

The author would like to acknowledge National Natural Science Foundation of China [51909037] and State Key Laboratory of Clean Energy Utilization [ZJU-CEU

secondary impingement region (called Region II). Furthermore, high ambient temperature exerts more influence on the fuel film formation for the Region II.

*DOI: http://dx.doi.org/10.5772/intechopen.95848*

of the fuel are the main reasons for it.

and area under both *T*amb = 298 K and *T*amb = 433 K.

it extends with an increase in ambient pressure.

CFD Computational fluid dynamics DISI Direct injection spark ignition RIM Refractive index matching

PFI Port fuel injection UHC Unburned hydrocarbon PM Particulate matter PN Particle number EV Electric vehicle

*T*sat Saturated temperatures fps Frames per second *S* Spray tip penetration *H*<sup>i</sup> Impinging spray height *L*<sup>w</sup> Distance to wall *L*<sup>r</sup> Radial distance ASOI After start of injection EOI End of injection *T*amb Ambient temperature *P*inj Injection pressure *P*amb Ambient pressure

**Acknowledgements**

2019005].

**Nomenclature**

secondary impingement region (called Region II). Furthermore, high ambient temperature exerts more influence on the fuel film formation for the Region II.


### **Acknowledgements**

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

**48**

**4. Conclusions**

*Film evaporates from "hollow".*

**Figure 30.**

**Figure 29.**

*"Hollow" development.*

major conclusions are summarized as follows:

This chapter discussed about the fuel spray and impingement under non-evaporation and evaporation conditions. The impinging spray development and fuel film formation as well as evaporation characteristics were checked and compared, the

*Schematic conceptual model. (a) Film evaporates from periphery. (b) With evaporation, "hollow"occurs. (c)* 

1.Different impingement regions were characterized through the formation of fuel film, known as the primary impingement region (called Region I) and the

The author would like to acknowledge National Natural Science Foundation of China [51909037] and State Key Laboratory of Clean Energy Utilization [ZJU-CEU 2019005].

#### **Nomenclature**


*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

#### **Author details**

Hongliang Luo Hiroshima University, Higashi-Hiroshima, Japan Foshan University, Foshan, P.R. China

\*Address all correspondence to: luo@hiroshima-u.ac.jp

© 2021 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

**51**

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines*

Experiments in fluids 2004; 37 (5): 745-762. https://doi.org/10.1007/

[8] Andreassi L, Ubertini S, & Allocca L. Experimental and numerical analysis of high pressure diesel spray–wall interaction. International journal of multiphase flow 2007; 33(7): 742-765. https://doi.org/10.1016/j. ijmultiphaseflow.2007.01.003

[9] Guo M, Shimasaki N, Nishida K, Ogata Y, Wada Y. Experimental study on fuel spray characteristics under atmospheric and pressurized cross-flow conditions. Fuel 2016; 184:846-855. https://doi.org/10.1016/j.

[10] Luo H, Uchitomi S, Nishida K, Ogata Y, Zhang W, Fujikawa T. Experimental Investigation on Fuel Film Formation of Spray Impingement on Flat Walls with Different Surface Roughness. Atomization and Sprays, 2017; 27(7): 611-628. DOI: 10.1615/

s00348-004-0866-3

fuel.2016.07.083

AtomizSpr.2017019706

[11] Luo H, Nishida K, Ogata Y. Evaporation characteristics of fuel adhesion on the wall after spray

impingement under different conditions through RIM measurement system. Fuel, 2019;258: 116163. https://doi. org/10.1016/j.fuel.2019.116163

[12] Luo H, Nishida K, Uchitomi S, Ogata Y, Zhang W, Fujikawa T. Effect of Spray Impinging Distance on Piston Top Fuel Adhesion in Direct Injection Gasoline Engines. International Journal of Engine Research, 2020; 21(5):742-754. https://doi. org/10.1177/1468087418774175

[13] Luo H, Nishida K, Uchitomi S, Ogata Y, Zhang W, Fujikawa T. Effect of temperature on fuel adhesion under spray-wall impingement condition.

*DOI: http://dx.doi.org/10.5772/intechopen.95848*

[1] Maly RR. State of the art and future needs in SI engine combustion. Symposium (International) on Combustion 1994; 25(1): 111- 124. https://doi.org/10.1016/ S0082-0784(06)80635-1

[2] Drake MC and Haworth DC.

[3] Kalantari D, Tropea C. Spray impact onto flat and rigid walls: Empirical characterization and modelling. International Journal of Multiphase Flow 2007; 33(5): 525-544. https://doi.org/10.1016/j. ijmultiphaseflow.2006.09.008

[4] Tanaka D, Uchida R, Noda T, Kolbeck A, Henkel S, Hardalupas Y, Taylor A, Aradi, A. Effects of fuel properties associated with in-cylinder behavior on particulate number from a direct injection gasoline engine. SAE Technical Paper, 2017-01-1002; 2017. https://doi.org/10.4271/2017-01-1002

[5] Zhao F, Lai MC, Harrington DL. Automotive spark-ignited directinjection gasoline engines. Progress in energy and combustion science 1999; 25(5): 437-562. https://doi.org/10.1016/

S0360-1285(99)00004-0

pecs.2010.01.002

[6] Moreira ALN, Moita AS, & Panao MR. Advances and challenges in explaining fuel spray impingement: How much of single droplet impact research is useful?. Progress in energy and combustion science 2010; 36(5): 554-580. https://doi.org/10.1016/j.

[7] Park SW and Lee CS. Macroscopic and microscopic characteristics of a fuel spray impinged on the wall.

using optical diagnostics and numerical modeling. Proceedings of the Combustion Institute 2007; 31(1): 99-124. https://doi.org/10.1016/j.

proci.2006.08.120

Advanced gasoline engine development

**References**

*Experimental Investigations on Fuel Spray and Impingement for Gasoline Direct Injection Engines DOI: http://dx.doi.org/10.5772/intechopen.95848*

#### **References**

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

**50**

**Author details**

Hongliang Luo

P.R. China

Hiroshima University, Higashi-Hiroshima, Japan Foshan University, Foshan,

© 2021 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium,

\*Address all correspondence to: luo@hiroshima-u.ac.jp

provided the original work is properly cited.

[1] Maly RR. State of the art and future needs in SI engine combustion. Symposium (International) on Combustion 1994; 25(1): 111- 124. https://doi.org/10.1016/ S0082-0784(06)80635-1

[2] Drake MC and Haworth DC. Advanced gasoline engine development using optical diagnostics and numerical modeling. Proceedings of the Combustion Institute 2007; 31(1): 99-124. https://doi.org/10.1016/j. proci.2006.08.120

[3] Kalantari D, Tropea C. Spray impact onto flat and rigid walls: Empirical characterization and modelling. International Journal of Multiphase Flow 2007; 33(5): 525-544. https://doi.org/10.1016/j. ijmultiphaseflow.2006.09.008

[4] Tanaka D, Uchida R, Noda T, Kolbeck A, Henkel S, Hardalupas Y, Taylor A, Aradi, A. Effects of fuel properties associated with in-cylinder behavior on particulate number from a direct injection gasoline engine. SAE Technical Paper, 2017-01-1002; 2017. https://doi.org/10.4271/2017-01-1002

[5] Zhao F, Lai MC, Harrington DL. Automotive spark-ignited directinjection gasoline engines. Progress in energy and combustion science 1999; 25(5): 437-562. https://doi.org/10.1016/ S0360-1285(99)00004-0

[6] Moreira ALN, Moita AS, & Panao MR. Advances and challenges in explaining fuel spray impingement: How much of single droplet impact research is useful?. Progress in energy and combustion science 2010; 36(5): 554-580. https://doi.org/10.1016/j. pecs.2010.01.002

[7] Park SW and Lee CS. Macroscopic and microscopic characteristics of a fuel spray impinged on the wall.

Experiments in fluids 2004; 37 (5): 745-762. https://doi.org/10.1007/ s00348-004-0866-3

[8] Andreassi L, Ubertini S, & Allocca L. Experimental and numerical analysis of high pressure diesel spray–wall interaction. International journal of multiphase flow 2007; 33(7): 742-765. https://doi.org/10.1016/j. ijmultiphaseflow.2007.01.003

[9] Guo M, Shimasaki N, Nishida K, Ogata Y, Wada Y. Experimental study on fuel spray characteristics under atmospheric and pressurized cross-flow conditions. Fuel 2016; 184:846-855. https://doi.org/10.1016/j. fuel.2016.07.083

[10] Luo H, Uchitomi S, Nishida K, Ogata Y, Zhang W, Fujikawa T. Experimental Investigation on Fuel Film Formation of Spray Impingement on Flat Walls with Different Surface Roughness. Atomization and Sprays, 2017; 27(7): 611-628. DOI: 10.1615/ AtomizSpr.2017019706

[11] Luo H, Nishida K, Ogata Y. Evaporation characteristics of fuel adhesion on the wall after spray impingement under different conditions through RIM measurement system. Fuel, 2019;258: 116163. https://doi. org/10.1016/j.fuel.2019.116163

[12] Luo H, Nishida K, Uchitomi S, Ogata Y, Zhang W, Fujikawa T. Effect of Spray Impinging Distance on Piston Top Fuel Adhesion in Direct Injection Gasoline Engines. International Journal of Engine Research, 2020; 21(5):742-754. https://doi. org/10.1177/1468087418774175

[13] Luo H, Nishida K, Uchitomi S, Ogata Y, Zhang W, Fujikawa T. Effect of temperature on fuel adhesion under spray-wall impingement condition.

Fuel 2018, 234:56-65. https://doi. org/10.1016/j.fuel.2018.07.021

[14] Bai C, Gosman AD. Development of methodology for spray impingement simulation. J Engines 1995,104(3):550- 68. https://www.jstor.org/ stable/44633238

[15] Mundo CHR, Sommerfeld M, Tropea C. Droplet-wall collisions: experimental studies of the formation and breakup process. Int J Multiph Flow 1995;21(2):151-73.

[16] Ding CP, Sjöberg M, Vuilleumier D, Reuss DL, He X, Böhm B. Fuel film thickness measurements using refractive index matching in a stratified-charge SI engine operated on E30 and alkylate fuels. Exp Fluids 2018;59(3):59. https:// doi.org/10.1007/s00348-018-2512-5

[17] Maligne D, Bruneaux G. Time-resolved fuel film thickness measurement for direct injection SI engines using refractive index matching. SAE Tech Paper 2011. 2011-01-1215. https://doi.org/10.4271/2011-01-1215

[18] Bai C, Rusche H, Gosman AD. Modeling of gasoline spray impingement. Atom Sprays 2002;12 (1-3):1-27. DOI: 10.1615/AtomizSpr.v12. i123.10

[19] Baumgarten C. Mixture Formation in Internal Combustion Engines. Germany: Springer Verlag; 2006.

[20] Schulz F, Beyrau F. Systematic Investigation of Fuel Film Evaporation. SAE Technical Paper 2018-01-0310; 2018. https://doi. org/10.4271/2018-01-0310

[21] Kim H, Sung N. The effect of ambient pressure on the evaporation of a single droplet and a spray. Combust Flame 2003;135(3):261-70. https://doi. org/10.1016/S0010-2180(03)00165-2

[22] Tao M, Ge H, VanDerWege B, Zhao P. Fuel wall film effects on premixed flame propagation, quenching and emission. Int J Engine Res 2018. (Online First). https://doi. org/10.1177/1468087418799565

**Chapter 3**

**Abstract**

**1. Introduction**

transport.

**53**

*Kenedy Aliila Greyson*

Vehicles Power Consumption:

Agency (DART) in Tanzania

Case Study of Dar Rapid Transit

Energy consumption and its environmental impact are now among the most challenging problems in most developing cities. The common sources of energy used as the fuel in transportation sector include gasoline, diesel, natural gas, propane, biofuels, electricity, coal, and hydrogen. However, in Tanzania, diesel and gasoline are still the dominant source of energy used by public and private vehicles. We have experienced significant efforts of converting conventional vehicles (gasoline engines) to operate on Compressed Natural Gas (CNG) or on hybrid system (gasoline and natural gas) as an alternative source of energy in Tanzania. The CNG is considered as cleaner combustion energy used as a vehicular fuel alternative to gasoline or diesel. In this chapter, the amount of energy consumption from the fuel combustion, the impact of environmental health (toxicity gas emission), the cost of fuel used by the transit buses in terms of fuel energy consumption, and driving profile are discussed. The scope of this work is based on the total energy contained in the fuel only. The ability of the engine to transform the available energy from the fuel into useful work power (efficiency) is left to the designers and manufacturers.

**Keywords:** CNG, consumption, cost, diesel, emission, energy, gasoline

emission such as Compressed Natural Gas (CNG) and to develop electric

powertrain systems. The chapter concludes with a brief look at fuel consumption and carbon dioxide (CO2) emissions in public transport vehicles and other modes of

The Tanzania strategic plans to increase economy are focused on industrialization agenda. The implementation of this agenda will increase the energy demand, hence the heavy fuel consumptions in industries and transportations sector. Energy consumption and its environmental impact are now among the most challenging problems in the most cities. Dar es Salaam is one of those fastest growing cities in Tanzania with the population over 5 million and expected to be a mega city with more than 10 million by 2030. Traffic congestion is one of the most prevalent transport challenges in Dar es Salaam. Due to the traffic congestion there is an increase of fuel consumption and likely potential harmful emissions. Moreover, the high fuel consumption (the imported commodity) significantly affects the national economy. The strategies are to reduce the harmful emissions by using fuel with less

#### **Chapter 3**

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

[22] Tao M, Ge H, VanDerWege B, Zhao P. Fuel wall film effects on

and emission. Int J Engine Res 2018. (Online First). https://doi. org/10.1177/1468087418799565

premixed flame propagation, quenching

Fuel 2018, 234:56-65. https://doi. org/10.1016/j.fuel.2018.07.021

68. https://www.jstor.org/

stable/44633238

1995;21(2):151-73.

[16] Ding CP, Sjöberg M, Vuilleumier D, Reuss DL,

[17] Maligne D, Bruneaux G. Time-resolved fuel film thickness measurement for direct injection SI engines using refractive index matching. SAE Tech Paper 2011. 2011-01-1215. https://doi.org/10.4271/2011-01-1215

[18] Bai C, Rusche H, Gosman AD. Modeling of gasoline spray

impingement. Atom Sprays 2002;12 (1-3):1-27. DOI: 10.1615/AtomizSpr.v12.

[19] Baumgarten C. Mixture Formation in Internal Combustion Engines. Germany: Springer Verlag; 2006.

[20] Schulz F, Beyrau F. Systematic

Evaporation. SAE Technical Paper 2018-01-0310; 2018. https://doi. org/10.4271/2018-01-0310

[21] Kim H, Sung N. The effect of ambient pressure on the evaporation of a single droplet and a spray. Combust Flame 2003;135(3):261-70. https://doi. org/10.1016/S0010-2180(03)00165-2

Investigation of Fuel Film

[14] Bai C, Gosman AD. Development of methodology for spray impingement simulation. J Engines 1995,104(3):550-

[15] Mundo CHR, Sommerfeld M, Tropea C. Droplet-wall collisions: experimental studies of the formation and breakup process. Int J Multiph Flow

He X, Böhm B. Fuel film thickness measurements using refractive index matching in a stratified-charge SI engine operated on E30 and alkylate fuels. Exp Fluids 2018;59(3):59. https:// doi.org/10.1007/s00348-018-2512-5

**52**

i123.10

## Vehicles Power Consumption: Case Study of Dar Rapid Transit Agency (DART) in Tanzania

*Kenedy Aliila Greyson*

#### **Abstract**

Energy consumption and its environmental impact are now among the most challenging problems in most developing cities. The common sources of energy used as the fuel in transportation sector include gasoline, diesel, natural gas, propane, biofuels, electricity, coal, and hydrogen. However, in Tanzania, diesel and gasoline are still the dominant source of energy used by public and private vehicles. We have experienced significant efforts of converting conventional vehicles (gasoline engines) to operate on Compressed Natural Gas (CNG) or on hybrid system (gasoline and natural gas) as an alternative source of energy in Tanzania. The CNG is considered as cleaner combustion energy used as a vehicular fuel alternative to gasoline or diesel. In this chapter, the amount of energy consumption from the fuel combustion, the impact of environmental health (toxicity gas emission), the cost of fuel used by the transit buses in terms of fuel energy consumption, and driving profile are discussed. The scope of this work is based on the total energy contained in the fuel only. The ability of the engine to transform the available energy from the fuel into useful work power (efficiency) is left to the designers and manufacturers.

**Keywords:** CNG, consumption, cost, diesel, emission, energy, gasoline

#### **1. Introduction**

The Tanzania strategic plans to increase economy are focused on industrialization agenda. The implementation of this agenda will increase the energy demand, hence the heavy fuel consumptions in industries and transportations sector. Energy consumption and its environmental impact are now among the most challenging problems in the most cities. Dar es Salaam is one of those fastest growing cities in Tanzania with the population over 5 million and expected to be a mega city with more than 10 million by 2030. Traffic congestion is one of the most prevalent transport challenges in Dar es Salaam. Due to the traffic congestion there is an increase of fuel consumption and likely potential harmful emissions. Moreover, the high fuel consumption (the imported commodity) significantly affects the national economy. The strategies are to reduce the harmful emissions by using fuel with less emission such as Compressed Natural Gas (CNG) and to develop electric powertrain systems. The chapter concludes with a brief look at fuel consumption and carbon dioxide (CO2) emissions in public transport vehicles and other modes of transport.

The Government of Tanzania through Dar es Salaam Rapid Transit (DART) agency operates a Bus Rapid Transit (BRT) system in Dar es Salaam. BRT system aimed to be a cost effective sustainable transportation system for the city of Dar es Salaam to ensure fast and orderly flow of traffic on urban streets and roads. It is reported by DART agency that, in phase one project each trunk bus and feeder bus daily operate nearly 297 km and 245 km, respectively. However, BRT's bus fare is still higher than other operators (daladala) due to the high operation costs. DART projects are implemented in phases. According the DART Agency [1], the phase one of DART corridor mainly involved 20.9 km of roads: Kimara to Kivukoni (15.8 km), Magomeni to Morocco (3.4 km) and Fire to Kariakoo (1.7 km). Other facilities include 27 bus stations and 5 terminals. The phase two is along Kilwa Road from city centre to Mbagala Rangi Tatu; other parts of the corridor start from South Kawawa road at Morogoro Road junction to Mgulani JKT round with 20.3 km. The implementation of the phase three DART project goes through the Nyerere highways from Gongolamboto to the City Center and some area of the Uhuru Road from Tazara to Kariakoo to Gerezani, which covers a total of 23.6 km.

CNG to the industries outside the downstream pipeline, and to other customers for

*Vehicles Power Consumption: Case Study of Dar Rapid Transit Agency (DART) in Tanzania*

Since, transit buses are among the cost effective forms of mass transit; most cities consider transit buses as the backbone of the transport system. However, the environmental health impact and running cost must be closed managed. In this chapter we investigate the current situation of BRT system buses using diesel, CNG and gasoline engines. The fuel consumption which is mentioned as the factor that affects the running costs, the cost of the fuel and the amount of fuel per distance, driving profile, and the emitted gases (mainly carbon dioxide) are discussed.

In this section, we briefly consider the main energy sources; gasoline, diesel and

CNG as the input power to the IC engine vehicles. The fuel consumption, and overall efficiency, on the baseline scenario are discussed. Although urban driving is characterized by lower driving speeds and presumably more frequent stops which lead to higher consumption values [7, 8], this study is characterized by the special lanes (that is, not shared by other vehicles) for DART's buses. In this section, the energy from the CNG, gasoline and diesel fuels are analyzed where the difference in (potential) energy of reactants (exothermic reaction) and products are presented.

The heat is given off when a specified amount (say, one mole) of a methane (reactant) burns in oxygen gas (heat of combustion) during the reaction. The general bond dissociation energy equation (enthalpy change) for the reaction at a specified temperature through the reaction is converted into a motive force for the vehicle is in

shown in Eq. (1) [9]. The heating of the fossil fuel results in CO2 emission.

ð Þ*<sup>g</sup>* ! *CO*<sup>2</sup>

*<sup>O</sup>*<sup>2</sup> ! *xCO*<sup>2</sup> <sup>þ</sup> *<sup>y</sup>*

where *CxHy* is the generic chemical formula for the fossil fuel, and *O2*, the oxygen in the input, and the output of the reaction are the emitted carbon dioxide

Methane (natural gas), which is a fossil fuel burns in oxygen to release energy as shown in Eq. (2). From **Table 1**, the combustion of methane gas will release 802.3 KJ/mol which is equivalent to calorific values of 50.1 KJ/g. The energy content of different fuels is made up of different chemical compounds, that is, bond breaking

2

*H*2*O* þ Energy (1)

ð Þ*<sup>g</sup>* <sup>þ</sup> <sup>2</sup>*H*2*O*ð Þ*<sup>g</sup>* <sup>þ</sup> Energy (2)

4

*CxHy* <sup>þ</sup> *<sup>x</sup>* <sup>þ</sup> *<sup>y</sup>*

(*CO2*), water (*H2O*) and the released energy.

energies and bond forming energies reaction.

ð Þ*<sup>g</sup>* <sup>þ</sup> <sup>2</sup>*O*<sup>2</sup>

*CH*<sup>4</sup>

Since both the gas and petrol engines run according to the Otto principle with a spark plug ignition, there are several centers in Tanzania converting petrol engines to work in hybrid systems. Dar es Salaam Institute of Technology (DIT) and University of Dar es Salaam (UDSM) are among these centers. It is therefore very important that the plans for the natural gases infrastructure expansions get real so that suppliers and operators can invest in proper natural gases products to be used effectively in transportation, industry and domestic sectors. Currently, at CNG stations, gas flow to the vehicle is measured by mass for sale with dispensers designed with compensation for temperature variations to ensure accurate

heating and cooking purposes.

*DOI: http://dx.doi.org/10.5772/intechopen.99031*

quantities are delivered to the tank [6].

**2. Propulsion energy**

**2.1 Natural gas**

**55**

Motor vehicles have been powered by gasoline, diesel, steam, gas turbine, and electric. [2]. The common Internal Combustion (IC) engines are gasoline sparkignition and diesel compression-ignition engines involving the combustion of a fuel inside a chamber that results in the expansion of the air/fuel mixture to produce mechanical work. In Tanzania, gasoline is the most common ground-transportation fuel, mostly used by private vehicles, followed by diesel used by most public vehicles and heavy duty trucks. As it is for the most public buses, BRT buses also use diesel as a fossil fuel. Diesel fuel in compression ignition engines produces a high level of toxicity in emission gases which leads to a health and environmental hazard [3]. The emitted gases associated with diesel fuel include nitrous oxides (NOx), carbon monoxide (CO), and carbon dioxide (CO2).

Tanzania has huge reserves of Natural Gas that city buses, especially the BRT systems, can use so as to minimize the harmful emissions. The development of gas supply infrastructures in Dar es Salam, Mtwara, and Coast Region are the examples of the ongoing initiatives. Recently, the natural gas, CNG in particular, has been considered to be a potential replacement and alternative to diesel and petrol fuels in vehicles due to its lower hazardous emission of gases in the environment. Applications of natural gas have been given much attention among stakeholders and researchers due to the remarkable attributes towards greener transportation. Studies on the use of natural gas in trains, buses, trucks, motorcycles, scooters and bicycles have been presented in various research reports. Since Tanzania has huge reserves of natural gas, this market growth increases with the users in various sectors including transportation.

Natural gas is used in transportation sector (public and private vehicles), industries, and domestic. If the use of natural gas resource is adopted effectively, the life of the people will improve both economically and environmentally. Therefore, a well-designed plan for the use of the natural gas in transmission and distribution is important. The objective is to optimize the utilization of gas resource as the alternative fuel energy in various sectors and the environmental safety. The strategy is to emphasize on the supply of CNG for transport sector use, and Piped Natural Gas (PNG) for domestic, commercial and industrial sectors in Tanzania. The CNG is mostly used in a public transit transportation fuel in other countries. CNG is a processed fluid gas into a high-pressure natural gas compression in a tube [4]. Compared to gasoline and diesel, vehicles powered by CNG emit less carbon monoxide, nitrogen oxides (NOx), and particulate matter [5]. Apart from the government initiatives of using CNG in transportation, there are strategic plans to supply

*Vehicles Power Consumption: Case Study of Dar Rapid Transit Agency (DART) in Tanzania DOI: http://dx.doi.org/10.5772/intechopen.99031*

CNG to the industries outside the downstream pipeline, and to other customers for heating and cooking purposes.

Since both the gas and petrol engines run according to the Otto principle with a spark plug ignition, there are several centers in Tanzania converting petrol engines to work in hybrid systems. Dar es Salaam Institute of Technology (DIT) and University of Dar es Salaam (UDSM) are among these centers. It is therefore very important that the plans for the natural gases infrastructure expansions get real so that suppliers and operators can invest in proper natural gases products to be used effectively in transportation, industry and domestic sectors. Currently, at CNG stations, gas flow to the vehicle is measured by mass for sale with dispensers designed with compensation for temperature variations to ensure accurate quantities are delivered to the tank [6].

Since, transit buses are among the cost effective forms of mass transit; most cities consider transit buses as the backbone of the transport system. However, the environmental health impact and running cost must be closed managed. In this chapter we investigate the current situation of BRT system buses using diesel, CNG and gasoline engines. The fuel consumption which is mentioned as the factor that affects the running costs, the cost of the fuel and the amount of fuel per distance, driving profile, and the emitted gases (mainly carbon dioxide) are discussed.

#### **2. Propulsion energy**

The Government of Tanzania through Dar es Salaam Rapid Transit (DART)

Motor vehicles have been powered by gasoline, diesel, steam, gas turbine, and electric. [2]. The common Internal Combustion (IC) engines are gasoline sparkignition and diesel compression-ignition engines involving the combustion of a fuel inside a chamber that results in the expansion of the air/fuel mixture to produce mechanical work. In Tanzania, gasoline is the most common ground-transportation fuel, mostly used by private vehicles, followed by diesel used by most public vehicles and heavy duty trucks. As it is for the most public buses, BRT buses also use diesel as a fossil fuel. Diesel fuel in compression ignition engines produces a high level of toxicity in emission gases which leads to a health and environmental hazard [3]. The emitted gases associated with diesel fuel include nitrous oxides (NOx),

Tanzania has huge reserves of Natural Gas that city buses, especially the BRT systems, can use so as to minimize the harmful emissions. The development of gas supply infrastructures in Dar es Salam, Mtwara, and Coast Region are the examples of the ongoing initiatives. Recently, the natural gas, CNG in particular, has been considered to be a potential replacement and alternative to diesel and petrol fuels in vehicles due to its lower hazardous emission of gases in the environment. Applications of natural gas have been given much attention among stakeholders and researchers due to the remarkable attributes towards greener transportation. Studies on the use of natural gas in trains, buses, trucks, motorcycles, scooters and bicycles have been presented in various research reports. Since Tanzania has huge reserves of natural gas, this market growth increases with the users in various

Natural gas is used in transportation sector (public and private vehicles), industries, and domestic. If the use of natural gas resource is adopted effectively, the life of the people will improve both economically and environmentally. Therefore, a well-designed plan for the use of the natural gas in transmission and distribution is important. The objective is to optimize the utilization of gas resource as the alternative fuel energy in various sectors and the environmental safety. The strategy is to emphasize on the supply of CNG for transport sector use, and Piped Natural Gas (PNG) for domestic, commercial and industrial sectors in Tanzania. The CNG is mostly used in a public transit transportation fuel in other countries. CNG is a processed fluid gas into a high-pressure natural gas compression in a tube [4]. Compared to gasoline and diesel, vehicles powered by CNG emit less carbon monoxide, nitrogen oxides (NOx), and particulate matter [5]. Apart from the government initiatives of using CNG in transportation, there are strategic plans to supply

carbon monoxide (CO), and carbon dioxide (CO2).

sectors including transportation.

**54**

agency operates a Bus Rapid Transit (BRT) system in Dar es Salaam. BRT system aimed to be a cost effective sustainable transportation system for the city of Dar es Salaam to ensure fast and orderly flow of traffic on urban streets and roads. It is reported by DART agency that, in phase one project each trunk bus and feeder bus daily operate nearly 297 km and 245 km, respectively. However, BRT's bus fare is still higher than other operators (daladala) due to the high operation costs. DART projects are implemented in phases. According the DART Agency [1], the phase one of DART corridor mainly involved 20.9 km of roads: Kimara to Kivukoni (15.8 km), Magomeni to Morocco (3.4 km) and Fire to Kariakoo (1.7 km). Other facilities include 27 bus stations and 5 terminals. The phase two is along Kilwa Road from city centre to Mbagala Rangi Tatu; other parts of the corridor start from South Kawawa road at Morogoro Road junction to Mgulani JKT round with 20.3 km. The implementation of the phase three DART project goes through the Nyerere highways from Gongolamboto to the City Center and some area of the Uhuru Road from Tazara to Kariakoo to Gerezani, which covers a total of

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

23.6 km.

In this section, we briefly consider the main energy sources; gasoline, diesel and CNG as the input power to the IC engine vehicles. The fuel consumption, and overall efficiency, on the baseline scenario are discussed. Although urban driving is characterized by lower driving speeds and presumably more frequent stops which lead to higher consumption values [7, 8], this study is characterized by the special lanes (that is, not shared by other vehicles) for DART's buses. In this section, the energy from the CNG, gasoline and diesel fuels are analyzed where the difference in (potential) energy of reactants (exothermic reaction) and products are presented.

#### **2.1 Natural gas**

The heat is given off when a specified amount (say, one mole) of a methane (reactant) burns in oxygen gas (heat of combustion) during the reaction. The general bond dissociation energy equation (enthalpy change) for the reaction at a specified temperature through the reaction is converted into a motive force for the vehicle is in shown in Eq. (1) [9]. The heating of the fossil fuel results in CO2 emission.

$$\rm C\_xH\_y + \left(x + \frac{y}{4}\right)O\_2 \rightarrow xCO\_2 + \frac{y}{2}H\_2O + \text{Energy} \tag{1}$$

where *CxHy* is the generic chemical formula for the fossil fuel, and *O2*, the oxygen in the input, and the output of the reaction are the emitted carbon dioxide (*CO2*), water (*H2O*) and the released energy.

Methane (natural gas), which is a fossil fuel burns in oxygen to release energy as shown in Eq. (2). From **Table 1**, the combustion of methane gas will release 802.3 KJ/mol which is equivalent to calorific values of 50.1 KJ/g. The energy content of different fuels is made up of different chemical compounds, that is, bond breaking energies and bond forming energies reaction.

$$\text{CH}\_4\text{(g)} + \text{2O}\_2\text{(g)} \rightarrow \text{CO}\_2\text{(g)} + \text{2H}\_2\text{O}^{\text{(g)}} + \text{Energy} \tag{2}$$

#### *Internal Combustion Engine Technology and Applications of Biodiesel Fuel*


**Table 1.**

*Average bond energies of common bonds.*

#### **2.2 Gasoline**

Energy from the gasoline (petrol) released during the reaction is as shown in Eq. (1). From the general reaction, the octane (reactant) burns in oxygen gas to release energy from the reaction. The bond dissociation energy (enthalpy change) for the reaction at a specified temperature through the gasoline chemical reaction is shown in Eq. (3).

$$\text{2C}\_8\text{H}\_{18}^{(l)} + \text{25O}\_2^{(g)} \rightarrow \text{16CO}\_2^{(g)} + \text{18H}\_2\text{O}^{(g)} + \text{Energy} \tag{3}$$

power provides the overall efficiency of the IC engine. It should be noted that, modern engines (petrol and diesel) have better efficiency. **Figure 1** shows the forces acting on the moving bus. Therefore, the discussion assumes the better

*Vehicles Power Consumption: Case Study of Dar Rapid Transit Agency (DART) in Tanzania*

*Fd* <sup>¼</sup> <sup>1</sup> 2

),

),

By definition, the rolling resistance is the combination of all frictional load forces of the tire on the road surface and the friction within the BRT bus. Aerodynamic drag is the resistance of air to the movement of the vehicle. The gradability of the vehicle load power can increase or decrease depending on whether the car is

It has been mentioned that BRT buses utilizes diesel as a fossil fuel. In analyzing the power consumed by the BRT bus, we need to understand the components of forces and resistance (opposing forces) that act on it. These forces and resistance on the bus are; rolling resistance due to tire and road interaction, aerodynamic drag resistance, grade resistance depends on motion towards the up-hill or the down-hill, and accelerating force due to the motion of the vehicle mass as shown in Eq. (5)-(8)

*Fr* ¼ *crmg* cos *α* (5)

*Fg* ¼ *mg* sin *α* (7) *Fa* ¼ *ma* (8)

> /s)<sup>2</sup> ),

> > ).

*ρcdAv*<sup>2</sup> (6)

performance of engines.

*Forces acting on a moving BRT bus.*

[10–13].

**Figure 1.**

where,

**57**

*Fr* is the rolling resistance,

*Fg* is the grade resistance, *Fa* is accelerating force,

*m* is the vehicle mass (N),

*Fd* is the aerodynamic drag resistance,

*cr* is the rolling resistance coefficient,

*cd* is the wind resistance coefficient (N/(m<sup>2</sup>

*g* is the gravitational constant (nominally 9.81 m/s<sup>2</sup>

*α* is the road grade angle (radian),

*A* is the vehicle frontal area (m<sup>2</sup>

*v* is the vehicle speed (m/s), *a* is the vehicle acceleration (m/s<sup>2</sup>

*ρ* density of air, and.

**3.1 Internal combustion engine power**

*DOI: http://dx.doi.org/10.5772/intechopen.99031*

The calculation reveals that the calorific value of the gasoline (petrol) is approximately 42 KJ/g.

#### **2.3 Diesel**

Energy from diesel is released during the reaction shown in Eq. (1) where the dodecane reactant, in the case of diesel, burns in oxygen gas to release energy from the reaction. The bond breaking energies and bond forming energies reaction is shown in Eq. (4).

$$4\text{C}\_{12}\text{H}\_{23}^{(l)} + 7\text{1O}\_{2}\text{(g)} \rightarrow 12\text{CO}\_{2}\text{(g)} + 4\text{6H}\_{2}\text{O}^{(g)} + \text{Energy} \tag{4}$$

The calculation reveals that the calorific value of the diesel is approximately 43 KJ/g.

#### **3. Longitudinal dynamics of the vehicle**

Vehicle dynamics depend on tire and road contact forces and torques, mass of the vehicle, road profile (road grade angles), ambient conditions, and driving profile. These parameters contribute into the total fuel consumption. The total power consumption is defines by the vehicle dynamics and distance travelled by the vehicle. The IC engine vehicles transform the chemical energy (from gasoline/ petrol, diesel, or CNG) into useful work (power consumed). The energy released by the gasoline, diesel, and CNG driven vehicles is discussed here. In fact, there are various factors affecting energy consumption when using IC engines. The ratio of power transmitted in a rotating shaft, referred to as brake power, and engine input

*Vehicles Power Consumption: Case Study of Dar Rapid Transit Agency (DART) in Tanzania DOI: http://dx.doi.org/10.5772/intechopen.99031*

**Figure 1.** *Forces acting on a moving BRT bus.*

**2.2 Gasoline**

*Average bond energies of common bonds.*

**Table 1.**

shown in Eq. (3).

**2.3 Diesel**

43 KJ/g.

**56**

shown in Eq. (4).

approximately 42 KJ/g.

<sup>2</sup>*C*8*H*18ð Þ*<sup>l</sup>* <sup>þ</sup> <sup>25</sup>*O*<sup>2</sup>

<sup>4</sup>*C*12*H*23ð Þ*<sup>l</sup>* <sup>þ</sup> <sup>71</sup>*O*<sup>2</sup>

**3. Longitudinal dynamics of the vehicle**

Energy from the gasoline (petrol) released during the reaction is as shown in Eq. (1). From the general reaction, the octane (reactant) burns in oxygen gas to release energy from the reaction. The bond dissociation energy (enthalpy change) for the reaction at a specified temperature through the gasoline chemical reaction is

**Bond Type Energy (kJ/mole)** C–H 414 C–C 347 C=C 615 O–H 463 C–O 360 C=O 728 H–H 436 O=O 498

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

ð Þ*<sup>g</sup>* <sup>þ</sup> <sup>18</sup>*H*2*O*ð Þ*<sup>g</sup>* <sup>þ</sup> Energy (3)

ð Þ*<sup>g</sup>* <sup>þ</sup> <sup>46</sup>*H*2*O*ð Þ*<sup>g</sup>* <sup>þ</sup> Energy (4)

ð Þ*<sup>g</sup>* ! <sup>16</sup>*CO*<sup>2</sup>

The calculation reveals that the calorific value of the gasoline (petrol) is

ð Þ*<sup>g</sup>* ! <sup>12</sup>*CO*<sup>2</sup>

The calculation reveals that the calorific value of the diesel is approximately

Vehicle dynamics depend on tire and road contact forces and torques, mass of the vehicle, road profile (road grade angles), ambient conditions, and driving profile. These parameters contribute into the total fuel consumption. The total power consumption is defines by the vehicle dynamics and distance travelled by the vehicle. The IC engine vehicles transform the chemical energy (from gasoline/ petrol, diesel, or CNG) into useful work (power consumed). The energy released by the gasoline, diesel, and CNG driven vehicles is discussed here. In fact, there are various factors affecting energy consumption when using IC engines. The ratio of power transmitted in a rotating shaft, referred to as brake power, and engine input

Energy from diesel is released during the reaction shown in Eq. (1) where the dodecane reactant, in the case of diesel, burns in oxygen gas to release energy from the reaction. The bond breaking energies and bond forming energies reaction is

power provides the overall efficiency of the IC engine. It should be noted that, modern engines (petrol and diesel) have better efficiency. **Figure 1** shows the forces acting on the moving bus. Therefore, the discussion assumes the better performance of engines.

#### **3.1 Internal combustion engine power**

It has been mentioned that BRT buses utilizes diesel as a fossil fuel. In analyzing the power consumed by the BRT bus, we need to understand the components of forces and resistance (opposing forces) that act on it. These forces and resistance on the bus are; rolling resistance due to tire and road interaction, aerodynamic drag resistance, grade resistance depends on motion towards the up-hill or the down-hill, and accelerating force due to the motion of the vehicle mass as shown in Eq. (5)-(8) [10–13].

$$F\_r = c\_r m \text{g } \cos a \tag{5}$$

$$F\_d = \frac{1}{2}\rho c\_d A \nu^2 \tag{6}$$

$$F\_{\mathfrak{g}} = \mathfrak{mg}\sin a \tag{7}$$

$$F\_a = ma\tag{8}$$

where,

*Fr* is the rolling resistance, *Fd* is the aerodynamic drag resistance, *Fg* is the grade resistance, *Fa* is accelerating force, *cr* is the rolling resistance coefficient, *m* is the vehicle mass (N), *α* is the road grade angle (radian), *cd* is the wind resistance coefficient (N/(m<sup>2</sup> /s)<sup>2</sup> ), *A* is the vehicle frontal area (m<sup>2</sup> ), *v* is the vehicle speed (m/s), *a* is the vehicle acceleration (m/s<sup>2</sup> ), *ρ* density of air, and.

*g* is the gravitational constant (nominally 9.81 m/s<sup>2</sup> ).

By definition, the rolling resistance is the combination of all frictional load forces of the tire on the road surface and the friction within the BRT bus. Aerodynamic drag is the resistance of air to the movement of the vehicle. The gradability of the vehicle load power can increase or decrease depending on whether the car is

ascending or descending an incline. Therefore, the vehicle power needed are typically based on vehicle acceleration requirements, usually specified as the time to accelerate and depends on the maximum available torque and maximum available power of the propulsion system. Then, the prime mover force which is the total tractive force, *FT for* the vehicle is given in Eq. (9).

$$F\_T = F\_r + F\_d + F\_a + F\_g \tag{9}$$

The tractive power and total tractive energy are given in (10) and (11), respectively.

$$P = F\_T \upsilon \tag{10}$$

stations and terminals are presented in **Table 3**. The total distance covered by the route code 001 is 15.6 km. The estimated energy consumption is calculated by using Eq. (11) based on the road profiles (grade angles at each sample) and driving profiles (speed, stops, accelerating, and braking). Then, the estimated energy consumption is converted into kWh to be compared with the available energy in the fossil fuels [14]. This consumption depends on the motion of the vehicle and how it behaves in motion. Due to the fact that, these dynamic parameters are complex and varies from different inputs, the discussion is focused only on the energy released

**SN Route Code Source Destination** 001 Kimara Kivukoni 002 Ubungo Kivukoni 003 Morocco Kivukoni 004 Kimara Gerezani 005 Ubungo Gerezani 006 Morocco Gerezani 007 Kimara Morocco 008 Ubungo Morocco 080 Muhimbili Gerezani 090 Kimara Mbezi

*Vehicles Power Consumption: Case Study of Dar Rapid Transit Agency (DART) in Tanzania*

Measurement of the driving profile can be done in different ways. The rotation

vector sensor and the gravity sensor are the most frequently used sensors for motion detection and monitoring. Speed, orientation and position of the DART's

from the fuel as the input of the engines.

*DART project phase one route codes and destinations.*

*DART phase one network Source: DART agency [1].*

*DOI: http://dx.doi.org/10.5772/intechopen.99031*

**4.1 Driving profile measurements**

**Figure 3.**

**Table 2.**

**59**

$$E = Pt = P\frac{s}{v} \tag{11}$$

where, *t* is the time spent, *s* is the distance travelled, *P* is the tractive power, and. E is the tractive energy.

#### **4. BRT bus power consumption**

The BRT system scenario and power consumption are presented in this section. The type of buses and routes are depicted in **Figures 2** and **3**, respectively. Our discussion of the vehicle's engine performance will consider the DART phase one network (route), from Kivukoni terminal to Kimara terminal with 21 stations and 3 terminals. The DART's buses in this route start from Kivukoni station to Kimara in Dar es Salaam as a one-segment of the route. For analysis purpose, diesel fuel, gasoline fuel, and CNG fuel are considered in the discussion section. Technically, the speed of the vehicle, *v* (km/h) is obtained to estimate the speed of the engine, *ω* (rpm) and torque, *τ* (Nm).

**Table 2** presents the route codes and destinations of the BRT buses in DART project phase one. The route code 001, from Kivukoni to Kimara, is used in this discussion. Other routes can be analyzed in the approach. Distances between

**Figure 2.** *BRT bus at the station.*

*Vehicles Power Consumption: Case Study of Dar Rapid Transit Agency (DART) in Tanzania DOI: http://dx.doi.org/10.5772/intechopen.99031*

**Figure 3.** *DART phase one network Source: DART agency [1].*


#### **Table 2.**

ascending or descending an incline. Therefore, the vehicle power needed are typically based on vehicle acceleration requirements, usually specified as the time to accelerate and depends on the maximum available torque and maximum available power of the propulsion system. Then, the prime mover force which is the total

The tractive power and total tractive energy are given in (10) and (11),

*<sup>E</sup>* <sup>¼</sup> *Pt* <sup>¼</sup> *<sup>P</sup> <sup>s</sup>*

The BRT system scenario and power consumption are presented in this section. The type of buses and routes are depicted in **Figures 2** and **3**, respectively. Our discussion of the vehicle's engine performance will consider the DART phase one network (route), from Kivukoni terminal to Kimara terminal with 21 stations and 3 terminals. The DART's buses in this route start from Kivukoni station to Kimara in Dar es Salaam as a one-segment of the route. For analysis purpose, diesel fuel, gasoline fuel, and CNG fuel are considered in the discussion section. Technically, the speed of the vehicle, *v* (km/h) is obtained to estimate the speed of the engine, *ω*

**Table 2** presents the route codes and destinations of the BRT buses in DART project phase one. The route code 001, from Kivukoni to Kimara, is used in this discussion. Other routes can be analyzed in the approach. Distances between

*v*

*FT* ¼ *Fr* þ *Fd* þ *Fa* þ *Fg* (9)

*P* ¼ *FTv* (10)

(11)

tractive force, *FT for* the vehicle is given in Eq. (9).

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

respectively.

where,

*t* is the time spent, *s* is the distance travelled, *P* is the tractive power, and. E is the tractive energy.

(rpm) and torque, *τ* (Nm).

**Figure 2.**

**58**

*BRT bus at the station.*

**4. BRT bus power consumption**

*DART project phase one route codes and destinations.*

stations and terminals are presented in **Table 3**. The total distance covered by the route code 001 is 15.6 km. The estimated energy consumption is calculated by using Eq. (11) based on the road profiles (grade angles at each sample) and driving profiles (speed, stops, accelerating, and braking). Then, the estimated energy consumption is converted into kWh to be compared with the available energy in the fossil fuels [14]. This consumption depends on the motion of the vehicle and how it behaves in motion. Due to the fact that, these dynamic parameters are complex and varies from different inputs, the discussion is focused only on the energy released from the fuel as the input of the engines.

#### **4.1 Driving profile measurements**

Measurement of the driving profile can be done in different ways. The rotation vector sensor and the gravity sensor are the most frequently used sensors for motion detection and monitoring. Speed, orientation and position of the DART's

#### *Internal Combustion Engine Technology and Applications of Biodiesel Fuel*


**4.2 Road profile measurements**

*DOI: http://dx.doi.org/10.5772/intechopen.99031*

where,

time and updated at *k* + 1 sample.

the same regardless of the engine used.

**5. Results discussions**

auxiliary usage.

**61**

The road profile and vehicle's speed (discussed earlier) are measured to deter-

The road grade angle (even vehicle speed) can be obtained by this method. The algorithm using KALMAN filters combines the use of accelerometer sensor and gyroscope (also known as gyro) sensor readings to obtain the speed and road profile. The readings in all orientations (*x*, *y*, *z*) are modeled as shown in

Eqs. (12)-(14) where *xxi* is the reading along *x* dimension, *xyi* is the reading along *y* dimension, and *xzi* is the reading along *z* dimension of the respective sensor (accelerometer and gyroscope). The *t* is the time of reading at *k* sample.

*<sup>x</sup>* <sup>þ</sup> *<sup>ν</sup>*ð Þ*<sup>k</sup>*

*<sup>y</sup>* <sup>þ</sup> *<sup>v</sup>*ð Þ*<sup>k</sup>*

ð Þ*k <sup>y</sup>*<sup>1</sup> þ *v* ð Þ*k*

*vx*, *vy* and *vz* are sensor readings in *x, y,* and *z* orientations at *t* interval of time, and *gx*, *gy* and *gz* are the road grade or speed in *x, y,* and *z* orientations at *k* sample

Finally, the readings at each sample are utilized to compute the road profile.

The engine torque is created on the crankshaft by the cylinder pressure pushing on the piston during the power stroke. Therefore, the maximum torque depends on the pressure pushed during the power stroke. However, in practice, the torque versus speed characteristic of an IC engines are not as linear as electric machines. The typical gasoline engine operates at no more than 10:1 compression ratio while the typical diesel engine may operate with a compression ratio as high as 25:1. It is obvious that, the higher the compression ratio, the better the overall engine efficiency. However, the scope of this work is based on the total energy contained in the fuel only. The great deal of the energy produced by a combustion engine is wasted. In this work, it is assumed that the speed of the vehicle along the route is

The driving profile (time/speed) of the BRT bus for route 001 obtained is shown in **Figure 4**. The distance of 15.8 km was travelled in about 37 minutes. The average speed of this BRT bus is 26 kilometer per hour (kph). Despite the 23 expected stops at the stations and terminals, there are other stops at the road junctions, and zebra crossings. Moreover, the drivers try to observe the speed limit of 50 kph and 60 kph along the route 001 road. The power consumption is proportional to the modes of the vehicle: traction (when the prime mover force, *Ft* > 0), braking (*Ft* < 0), and coasting (*Ft* = 0). **Figure 5** shows the average power consumption for the BRT bus on route 001. The energy used to brake or slow down a vehicle in a conventional vehicle is dissipated as heat in the braking system and lost to the vehicle. An electric vehicle can be designed to regenerate the energy and store it on the vehicle for

*<sup>x</sup> t* (12)

*<sup>y</sup> t* (13)

*<sup>y</sup>*<sup>2</sup> *t* (14)

*g*ð Þ *<sup>k</sup>*þ<sup>1</sup> *<sup>x</sup>* <sup>¼</sup> *<sup>g</sup>*ð Þ*<sup>k</sup>*

*g*ð Þ *<sup>k</sup>*þ<sup>1</sup> *<sup>y</sup>* <sup>¼</sup> *<sup>g</sup>*ð Þ*<sup>k</sup>*

*g* ð Þ *k*þ1 *<sup>y</sup>*<sup>1</sup> ¼ *g*

mine power consumed by the vehicle. The quality of the road profile defines comfort ride, quality of load and passenger transportation as well as external excitation of the ground vehicle [19]. This subsection presents the road condition monitoring where rotation vector sensor and the gravity sensor are used.

*Vehicles Power Consumption: Case Study of Dar Rapid Transit Agency (DART) in Tanzania*

**Table 3.**

*DART Route 001 distances between stations/terminals.*

bus were recorded by the use of the Global Positioning System (GPS) and Inertial Navigation System (INS) where integration of Inertia Measurement Unit (IMU) and GPS (GPS/IMU) are used [15].

There are several techniques to obtain the speed of the moving vehicles. For example, algorithm to estimate vehicle speed from accelerometer data generated by an onboard smart-phone [16] has been proposed in various literatures. However, according to authors in [16], speed estimation by the integration of the mobilephone accelerometer data will not yield accurate results, since the accelerometer data in the direction of motion is not pure acceleration, but involves white noise, phone sensor bias, vibration, gravity component, and other effects. Similarly, the use of GPS to measure vehicle speed, on the other hand, depends on the number of visible GPS satellites at the recording time. Other factors include model of the mobile phone, whether condition, radio noise, concurrent usage, location and type of terrain, mobile network triangulation, Wi-Fi network location, and magnetic fields. [17] depending on the algorithm used. The speed of the vehicle can be estimated after analyzing the acceleration characteristic values in advancing direction and vertical direction as explained in [18].

*Vehicles Power Consumption: Case Study of Dar Rapid Transit Agency (DART) in Tanzania DOI: http://dx.doi.org/10.5772/intechopen.99031*

#### **4.2 Road profile measurements**

The road profile and vehicle's speed (discussed earlier) are measured to determine power consumed by the vehicle. The quality of the road profile defines comfort ride, quality of load and passenger transportation as well as external excitation of the ground vehicle [19]. This subsection presents the road condition monitoring where rotation vector sensor and the gravity sensor are used.

The road grade angle (even vehicle speed) can be obtained by this method. The algorithm using KALMAN filters combines the use of accelerometer sensor and gyroscope (also known as gyro) sensor readings to obtain the speed and road profile. The readings in all orientations (*x*, *y*, *z*) are modeled as shown in Eqs. (12)-(14) where *xxi* is the reading along *x* dimension, *xyi* is the reading along *y* dimension, and *xzi* is the reading along *z* dimension of the respective sensor (accelerometer and gyroscope). The *t* is the time of reading at *k* sample.

$$\mathbf{g}\_{\mathbf{x}}^{(k+1)} = \mathbf{g}\_{\mathbf{x}}^{(k)} + \boldsymbol{\nu}\_{\mathbf{x}}^{(k)} \mathbf{t} \tag{12}$$

$$\mathbf{g}\_{\mathbf{y}}^{(k+1)} = \mathbf{g}\_{\mathbf{y}}^{(k)} + \boldsymbol{\nu}\_{\mathbf{y}}^{(k)} \mathbf{t} \tag{13}$$

$$\mathbf{g}\_{\mathbf{y}1}^{(k+1)} = \mathbf{g}\_{\mathbf{y}1}^{(k)} + \boldsymbol{\nu}\_{\mathbf{y}2}^{(k)} \mathbf{t} \tag{14}$$

where,

*vx*, *vy* and *vz* are sensor readings in *x, y,* and *z* orientations at *t* interval of time, and *gx*, *gy* and *gz* are the road grade or speed in *x, y,* and *z* orientations at *k* sample time and updated at *k* + 1 sample.

Finally, the readings at each sample are utilized to compute the road profile.

#### **5. Results discussions**

The engine torque is created on the crankshaft by the cylinder pressure pushing on the piston during the power stroke. Therefore, the maximum torque depends on the pressure pushed during the power stroke. However, in practice, the torque versus speed characteristic of an IC engines are not as linear as electric machines.

The typical gasoline engine operates at no more than 10:1 compression ratio while the typical diesel engine may operate with a compression ratio as high as 25:1. It is obvious that, the higher the compression ratio, the better the overall engine efficiency. However, the scope of this work is based on the total energy contained in the fuel only. The great deal of the energy produced by a combustion engine is wasted. In this work, it is assumed that the speed of the vehicle along the route is the same regardless of the engine used.

The driving profile (time/speed) of the BRT bus for route 001 obtained is shown in **Figure 4**. The distance of 15.8 km was travelled in about 37 minutes. The average speed of this BRT bus is 26 kilometer per hour (kph). Despite the 23 expected stops at the stations and terminals, there are other stops at the road junctions, and zebra crossings. Moreover, the drivers try to observe the speed limit of 50 kph and 60 kph along the route 001 road. The power consumption is proportional to the modes of the vehicle: traction (when the prime mover force, *Ft* > 0), braking (*Ft* < 0), and coasting (*Ft* = 0). **Figure 5** shows the average power consumption for the BRT bus on route 001. The energy used to brake or slow down a vehicle in a conventional vehicle is dissipated as heat in the braking system and lost to the vehicle. An electric vehicle can be designed to regenerate the energy and store it on the vehicle for auxiliary usage.

bus were recorded by the use of the Global Positioning System (GPS) and Inertial Navigation System (INS) where integration of Inertia Measurement Unit (IMU)

**SN Name Category Distance (km)** Kivukoni Terminal 0.0 Posta Zamani Station 1.14 Nyerere Square Station 1.52 Kisutu Station 2.25 DIT Station 2.62 Fire Station 3.18 Jangwani Station 3.85 Magomeni Mapipa Station 4.88 Usalama Station 5.6 Mwembe chai Station 6.12 Kagera Station 6.6 Argentina Station 7.25 Bakrhesa Station 7.77 Manzese Station 8.4 Urafiki Station 9.0 Shekilango Station 9.8 Ubungo Terminal 10.5 Ubungo Maji Station 11.2 Kibo Station 12.2 Kona Station 12.8 Baruti Station 13.38 Bucha Station 13.86 Korogwe Station 14.5 Kimara Terminal 15.8

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

There are several techniques to obtain the speed of the moving vehicles. For example, algorithm to estimate vehicle speed from accelerometer data generated by an onboard smart-phone [16] has been proposed in various literatures. However, according to authors in [16], speed estimation by the integration of the mobilephone accelerometer data will not yield accurate results, since the accelerometer data in the direction of motion is not pure acceleration, but involves white noise, phone sensor bias, vibration, gravity component, and other effects. Similarly, the use of GPS to measure vehicle speed, on the other hand, depends on the number of visible GPS satellites at the recording time. Other factors include model of the mobile phone, whether condition, radio noise, concurrent usage, location and type of terrain, mobile network triangulation, Wi-Fi network location, and magnetic fields. [17] depending on the algorithm used. The speed of the vehicle can be estimated after analyzing the acceleration characteristic values in advancing direc-

and GPS (GPS/IMU) are used [15].

*DART Route 001 distances between stations/terminals.*

**Table 3.**

**60**

tion and vertical direction as explained in [18].

**Figure 4.** *Driving profile of the BRT bus on route 001.*

#### **Figure 5.**

*Average power consumption for the BRT bus on route 001.*

The total energy consumed when the given vehicle (parameters in **Table 4**) on route 001has been computed to be approximately 19.952 kWh. In fossil fuel environment, this energy can be obtained from gasoline, diesel, or CNG. The calorific values of these fuels (Energy/mass) can be in KJ/g or kWh/Kg. In converting calorific value, energy/mass, in KJ/g of the fuel into calorific value in kWh/ Kg, we divide the energy/mass (KJ/g) value by 3.6. Gasoline and diesel densities are presented in **Table 5** [20]. Using information from **Table 5**, other conversions

4-speed transmission ratios, *rt* — [2.393 1.450 1.000 0.677]

Wheel radius, *rr* m 0.381 Engine efficiency — 0.92 Tank capacity is about liter 475 Maximum engine torque — 1200 Nm Maximum engine power — 284 kW

**Parameter Unit Value**

*Vehicles Power Consumption: Case Study of Dar Rapid Transit Agency (DART) in Tanzania*

Rolling Resistance Coefficient, *Cr* — 0.01 Vehicle mass, *m* kg 11750 Passenger mass kg 5590 Vehicle total mass kg 17340 Frontal area, *A* m2 6.93 Acceleration due to gravity, *g* m/s<sup>2</sup> 9.8 Final-drive ratio, *rf* — 3.37

/s)<sup>2</sup> 0.599

In Tanzania mainland, the Energy and Water Utilities Regulatory Authority (EWURA) is responsible to publish cap prices for petroleum products. The wholesale and retail prices vary from time to time depending on price of the imported petroleum product imported. For example, According to the released wholesale and retail prices released by EWURA on Wednesday, 3rd February 2021, the retail

(kWh/Kg and kWh/L) are presented in **Table 6**.

**Figure 7.**

**Table 4.**

**63**

*Transit bus parameters.*

*Energy consumption for the BRT bus on route 001.*

*DOI: http://dx.doi.org/10.5772/intechopen.99031*

Wind resistance coefficient N/(m<sup>2</sup>

**Figure 6.** *Section of power consumption (0–250 second).*

**Figure 6** depicts the section of power consumption (0–250 seconds), showing the stopping (or braking) periods. The more power is consumed when the bus is in accelerating mode. As discussed earlier, the power consumption (and emission) depends on the vehicle design parameters, auxiliary devices, driving profiles, road profiles and conditions. **Figure 7** shows the amount of energy consumption for the BRT bus on route 001 obtained by using Eq. (11).

*Vehicles Power Consumption: Case Study of Dar Rapid Transit Agency (DART) in Tanzania DOI: http://dx.doi.org/10.5772/intechopen.99031*

**Figure 7.**

*Energy consumption for the BRT bus on route 001.*


#### **Table 4.**

*Transit bus parameters.*

The total energy consumed when the given vehicle (parameters in **Table 4**) on route 001has been computed to be approximately 19.952 kWh. In fossil fuel environment, this energy can be obtained from gasoline, diesel, or CNG. The calorific values of these fuels (Energy/mass) can be in KJ/g or kWh/Kg. In converting calorific value, energy/mass, in KJ/g of the fuel into calorific value in kWh/ Kg, we divide the energy/mass (KJ/g) value by 3.6. Gasoline and diesel densities are presented in **Table 5** [20]. Using information from **Table 5**, other conversions (kWh/Kg and kWh/L) are presented in **Table 6**.

In Tanzania mainland, the Energy and Water Utilities Regulatory Authority (EWURA) is responsible to publish cap prices for petroleum products. The wholesale and retail prices vary from time to time depending on price of the imported petroleum product imported. For example, According to the released wholesale and retail prices released by EWURA on Wednesday, 3rd February 2021, the retail

**Figure 6** depicts the section of power consumption (0–250 seconds), showing the stopping (or braking) periods. The more power is consumed when the bus is in accelerating mode. As discussed earlier, the power consumption (and emission) depends on the vehicle design parameters, auxiliary devices, driving profiles, road profiles and conditions. **Figure 7** shows the amount of energy consumption for the

BRT bus on route 001 obtained by using Eq. (11).

*Section of power consumption (0–250 second).*

**Figure 4.**

**Figure 5.**

**Figure 6.**

**62**

*Driving profile of the BRT bus on route 001.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

*Average power consumption for the BRT bus on route 001.*

#### *Internal Combustion Engine Technology and Applications of Biodiesel Fuel*


**6. Conclusion**

**Author details**

**65**

Kenedy Aliila Greyson

The concerns of the global warming and environmental pollution have led to

*Vehicles Power Consumption: Case Study of Dar Rapid Transit Agency (DART) in Tanzania*

converting conventional vehicles to operate on natural gas is a good option for BRT system. The operating cost is low due the low cost of CNG and the CO2 gas emission is low compared to other fuel types (diesel and gasoline). Since CNG is available locally, (not imported), the national agenda of industrialization is supported by the

more severe regulations on CO2 and other pollutant emissions. Therefore,

Dar es Salaam Institute of Technology (DIT), Dar es Salaam, Tanzania

© 2021 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium,

\*Address all correspondence to: kenedyaliila@yahoo.com

provided the original work is properly cited.

use of CNG as an alternative source of energy.

*DOI: http://dx.doi.org/10.5772/intechopen.99031*

#### **Table 5.**

*Specific gravity of motor fuel (Kg/L).*


#### **Table 6.**

*Other units for energy mass of the fuel.*


#### **Table 7.**

*Energy consumed and cost of fuel used in route 001.*


#### **Table 8.**

*CO2 emission in one segment of Kivukoni-Kimara route.*

prices were as indicated in the underneath of **Table 7**. From Eq. (11) the average required energy for the BRT bus on route 001 approximately 19.952 kWh. This energy can be delivered by the gasoline, diesel or CNG as shown in **Table 7**. Therefore, with other factors assumed the same and perfect, gasoline IC engine bus would consume 2.32 liters of gasoline to cover the route while diesel IC engine would consume 1.9 liter of diesel in the same route. CNG engine bus, on the other hand, would 1.22 Kg of CNG gas. The total fuel cost required is shown in the extreme column of **Table 7**.

The emission of carbon dioxide (CO2) for an internal combustion engine to move a vehicle down the road depends on the type of fuel used. While 1 L of gasoline produces approximately 2.3 Kg of CO2, burning 1 L of diesel produces approximately 2.67 Kg of CO2. For the case of CNG, CO2 emission depends on the calorific gas (i.e. low calorific gas and high calorific gas). This emission ranges from 2.25 Kg of CO2 (low calorific gas) to 2.67 Kg of CO2 (high calorific gas) when burning 1 Kg of CNG. Hence, the total CO2 emission in route 001 is shown in **Table 8**. The CNG emission is still low compared to gasoline and diesel.

*Vehicles Power Consumption: Case Study of Dar Rapid Transit Agency (DART) in Tanzania DOI: http://dx.doi.org/10.5772/intechopen.99031*

#### **6. Conclusion**

The concerns of the global warming and environmental pollution have led to more severe regulations on CO2 and other pollutant emissions. Therefore, converting conventional vehicles to operate on natural gas is a good option for BRT system. The operating cost is low due the low cost of CNG and the CO2 gas emission is low compared to other fuel types (diesel and gasoline). Since CNG is available locally, (not imported), the national agenda of industrialization is supported by the use of CNG as an alternative source of energy.

#### **Author details**

prices were as indicated in the underneath of **Table 7**. From Eq. (11) the average required energy for the BRT bus on route 001 approximately 19.952 kWh. This energy can be delivered by the gasoline, diesel or CNG as shown in **Table 7**. Therefore, with other factors assumed the same and perfect, gasoline IC engine bus would consume 2.32 liters of gasoline to cover the route while diesel IC engine would consume 1.9 liter of diesel in the same route. CNG engine bus, on the other hand, would 1.22 Kg of CNG gas. The total fuel cost required is shown in the

**Fuel Kg/liter** Gasoline 0.737 Diesel 0.850

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

**Fuel KJ/g kWh/Kg kWh/L** Gasoline 42 11.67 8.6 Diesel 43 11.94 10.15 CNG 50.1 13.92 —

**Fuel Required energy (kWh) route 001 Fuel Cost\* Kg Liters** Gasoline — 2.32 4,377.84 Diesel — 1.9 3,475.10

**Fuel CO2 Emission (Kg)** Gasoline 5.336 Diesel 5.073 CNG 3.257\*

*Gasoline (petrol): 1887.00 TZS/Liter; Diesel 1829.00 TZS/Liter; CNG: 1550.00 TZS/kg.*

The emission of carbon dioxide (CO2) for an internal combustion engine to move a vehicle down the road depends on the type of fuel used. While 1 L of gasoline produces approximately 2.3 Kg of CO2, burning 1 L of diesel produces approximately 2.67 Kg of CO2. For the case of CNG, CO2 emission depends on the calorific gas (i.e. low calorific gas and high calorific gas). This emission ranges from 2.25 Kg of CO2 (low calorific gas) to 2.67 Kg of CO2 (high calorific gas) when burning 1 Kg of CNG. Hence, the total CO2 emission in route 001 is shown in **Table 8**. The CNG emission is still low compared to gasoline and diesel.

extreme column of **Table 7**.

*Emission when high calorific gas is used.*

**Table 5.**

**Table 6.**

*\**

*\**

**64**

**Table 8.**

**Table 7.**

*Specific gravity of motor fuel (Kg/L).*

*Other units for energy mass of the fuel.*

*Energy consumed and cost of fuel used in route 001.*

*CO2 emission in one segment of Kivukoni-Kimara route.*

Kenedy Aliila Greyson Dar es Salaam Institute of Technology (DIT), Dar es Salaam, Tanzania

\*Address all correspondence to: kenedyaliila@yahoo.com

© 2021 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

### **References**

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[3] Kojima M. Breathing Clean: Considering the switch to natural gas buses. World Bank Technical Paper No. 516, 2001; p.1-50, DOI: 10.1596/0-8213- 5040-4.

[4] Isworo P. Compressed natural gas technology for alternative fuel power plants," in Proc. 2nd International Conference on Energy, Environmental and Information System (ICENIS), August 14-15, 2018; Semarang, Indonesia. p. 1-4.

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[11] Wai K, Rong Y, Morris S. Simulation of a distance estimator for a battery electric vehicle. Alexandria Engineering Journal; 2015; 54 (3):359-371. DOI: 10.1016/j.aej.2015.04.008.

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[13] Zongxuan S, Guoming Z. Design and Control of Automotive Propulsion Systems. Boca Raton, Fla:CRC Press, 2015. DOI: 10.1201/b17947.

[14] Wai, K, Rong, Y, Morris S. Simulation of a distance estimator for a battery electric vehicle. Alexandria Engineering Journal. 2015; 54 (3): 359-371. DOI: 10.1016/j.aej.2015.04.008.

[15] Rogers M, Applied Mathematics in Integrated Navigation Systems. 3rd ed. *Vehicles Power Consumption: Case Study of Dar Rapid Transit Agency (DART) in Tanzania DOI: http://dx.doi.org/10.5772/intechopen.99031*

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[Accessed: 2020-07-12].

10.17226/12924.

Indonesia. p. 1-4.

02190-6.

5040-4.

Administration and Local Government, Dar Rapid Transit (DART) agency, Available from: https://www.dart.go.tz

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

[8] Braun A, Rid W. Energy consumption of an electric and an internal combustion passenger car. A comparative case study from real world data on the Erfurt circuit in Germany. In: Proceedings of 20th EURO Working Group on Transportation Meeting, (EWGT 2017); 4-6 September 2017; Budapest, Hungary. Transportation Research Procedia 2017; p. 468–475.

[9] John H, Abas Goodarzi G. Electric Powertrain: Energy Systems, Power Electronics and Drives for Hybrid, Electric and Fuel Cell Vehicles. Wiley; 2019. DOI: 10.1002/9781119063681.

[10] Sousa D, Costa J, Dente J. Electric

supercapacitors. In: Proceedings of the European Conference on Power Electronics and Applications; 2-5 September 2007; Aalborg, Denmark,

[11] Wai K, Rong Y, Morris S. Simulation of a distance estimator for a battery electric vehicle. Alexandria Engineering Journal; 2015; 54 (3):359-371. DOI:

[12] Łebkowski A. Studies of energy consumption by a city bus powered by a hybrid energy storage system in variable road conditions. In Energies. 2019; 12: 951. DOI: 10.3390/en12050951.

[13] Zongxuan S, Guoming Z. Design and Control of Automotive Propulsion Systems. Boca Raton, Fla:CRC Press,

Simulation of a distance estimator for a battery electric vehicle. Alexandria Engineering Journal. 2015; 54 (3): 359-371. DOI: 10.1016/j.aej.2015.04.008.

[15] Rogers M, Applied Mathematics in Integrated Navigation Systems. 3rd ed.

2015. DOI: 10.1201/b17947.

[14] Wai, K, Rong, Y, Morris S.

bicycle using batteries and

10.1016/j.aej.2015.04.008.

2007. p. 774 – 781.

[2] Committee on the Assessment of Technologies for Improving Light-Duty

Vehicle Fuel Economy, National Research Council. Assessment of fuel economy technologies for light-duty vehicles. Washington, DC: The National Academies Press, 2011. 14 p. DOI:

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Section 2

Application of Biodiesel

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Section 2
