**2. Design options**

As it can be deduced by the previous section, there are several options that can be explored in the design of 2-Stroke CI high speed engines. The most typical ones are listed below.


A synthetic comparison among the configurations is given in Table 1 while Figure 1 shows the relative layouts.


Supercharging further improves power density and fuel efficiency, as well as enhancing alti‐ tude performance. Diesel combustion allows a higher boosting level, in comparison to Spark Ignited engines, limited by knocking. In addition, high octane aviation gasoline is expected to be subject to strong limitations, due to its polluting emissions of lead, while a Diesel en‐ gine can burn a variety of fuels: besides automotive Diesel, also turbine fuels such as JP4 and JP5, and Jet A. Further advantages in comparison to gasoline power-plants are: reduced fire and explosion hazard, better in-flight reliability (no mixture control problems), no car‐ buretor icing problems and safe cabin heating from exhaust stacks (less danger of Carbon

As it can be deduced by the previous section, there are several options that can be explored in the design of 2-Stroke CI high speed engines. The most typical ones are listed below.

**1.** Uniflow scavenging with exhaust poppet valves and piston controlled inlet ports; exter‐

**2.** Uniflow scavenging with exhaust poppet valves and piston controlled inlet ports; exter‐ nal blower and 4-Stroke- like oil sump; indirect injection with a pre-chamber connected

**3.** Loop Scavenging with piston controlled transfer and exhaust ports; crankcase pump; indirect injection with a pre-chamber connected to the cylinder through one or more or‐

**4.** Uniflow scavenging with opposed pistons, twin crankshafts; external blower and oil sump; indirect injection with a pre-chamber connected to the cylinder through one or

**5.** Loop scavenging with inlet and exhaust poppet valves in the engine head; 4-Stroke-like crankcase and external blower; indirect injection with a pre-chamber connected to the

**6.** Loop scavenging with inlet and exhaust poppet valves in the engine head, 4-Stroke-like

**7.** Loop Scavenging with piston controlled transfer and exhaust ports; 4-Stroke-like crank‐ case and external blower; indirect injection with a pre-chamber connected to the cylin‐

**8.** Loop Scavenging with piston controlled transfer and exhaust ports; 4-Stroke-like crank‐ case and external blower; direct injection with a chamber carved in the engine head. A synthetic comparison among the configurations is given in Table 1 while Figure 1 shows

crankcase and external blower; direct injection, bowl in the piston.

nal blower and 4-Stroke- like oil sump; direct injection, bowl in the piston.

to the cylinder through one or more orifices.

cylinder through one or more orifices.

der through one or more orifices.

Monoxide intoxication).

152 Advances in Internal Combustion Engines and Fuel Technologies

**2. Design options**

ifices

more orifices

the relative layouts.

**Table 1.** Comparison among the different designs listed in the previous section. Grades: A=Excellent, B=Good, C=Average, D=Poor

From the scavenging quality point of view, uniflow scavenging is generally better than loop, even if the necessity of imparting a swirling motion to the inlet flow can spoil the advantage a little bit. Since the swirl requirement is more stringent for direct injection, DI Uniflow scav‐ enging configurations generally yield lower trapping and scavenging efficiency than Uni‐ flow IDI designs. Another advantage of the IDI design is the cost of the injection system, that can be of the mechanical type. The downsides are the low thermal efficiency and the limitation on power rating due to smoke emissions at high speed and load.

When scavenging is obtained only by means of piston controlled ports, the valve-train is absent. However, the advantage in terms of mechanical efficiency can be spoiled without a proper lubrication, or in the case of a double crankshaft (opposed piston design). Par‐ ticular care must be devoted when using a crankcase pump, since some oil uniformly dispersed in the airflow is generally not sufficient at high load. On the other hand, the combination of loop scavenging and crankcase pump enables a very compact design when power rating is low.

Except for the opposed pistons configuration, the piston-controlled ports design implies that a tumble motion is generated within the cylinder. The same type of flow field can be found in the designs with inlet and exhaust poppet valves, referred to as 5 and 6. The optimization of a DI combustion system without swirl is far from trivial and it requires a strong support by simulation and specific experiments, with ensuing rise of the engineering costs. The same problem may be faced in the development of an opposed piston design, because of the lack of reference in recent projects.

In general, every solution presented in table 1 has its own pros and cons, so that the best choice depends on the project specifics. In the authors' opinion, the most balanced solutions are #1 and #8.

losses), while minimizing short circuiting and the mixing between fresh charge and ex‐ haust gas. Another important issue is the conditioning of the mean in- cylinder flow field (swirl or tumble), which strongly affects both combustion and heat transfer. The op‐ timum intensity of the swirl/tumble rates depends on the type of combustion system, as well as on the specific project targets. As an example, the swirl ratio in DI engine with a bowl in the piston should be high enough to promote the diffusion of the fuel vapor in the chamber. However, an excessive mean turbulence is detrimental to spray penetration,

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The energy spent to pump the fresh charge across the cylinder is a fundamental parame‐ ter, even if not the only one, to assess the quality of the scavenging system. In order to find a simplified correlation among the average pressure drop across the cylinder (*Δp*) and the main engine parameters, the gas exchange process in a 2-Stroke engine can be idealized as a steady phenomenon, with the piston fixed at bottom dead center and both inlet and exhaust ports partially open, so that the geometric area of each port corre‐ sponds to the average effective area, calculated over the cycle. As a further simplifica‐ tion, the flow is assumed as uncompressible. According to these hypotheses, the mass

<sup>2</sup> , <sup>2</sup>

r

Combining equation 1 and 2, the following expression for *Δp* is found:

r

*p p A p eff av* r

Where ρ is the charge density, *DR* is the Delivery Ratio of the engine (ratio of the delivered fresh charge to the reference mass, calculated as the product of charge density to cylinder displacement), *U<sup>p</sup>* is the mean piston speed. *Aeff,av* is the average effective area of all the ports,

> 2 2 1

> > 2

, *p*

è ø

*A*

+

, 1 1 *T E Aeff av A A* =

Being *AT* the mean effective area of the transfer ports and *AE* the mean effective area of the

2 2

æ ö ç ÷ Dµ × × ×ç ÷

*p DR U Aeff av*

*P*

*DR A U*

×D = (1)

(2)

(3)

×××

and heat losses increase.

that can be expressed as:

exhaust ports.

flow rate across the cylinder can be expressed as:

**Figure 1.** Typical configurations of 2-Stroke CI high speed engines

#### **3. Scavenging systems**

The optimization of the scavenging process is one of the most challenging task in the de‐ sign of 2-Stroke engines. In fact, the geometry of the ports-cylinder assembly should be defined in order to guarantee a smooth path of the flow across the engine (low flow losses), while minimizing short circuiting and the mixing between fresh charge and ex‐ haust gas. Another important issue is the conditioning of the mean in- cylinder flow field (swirl or tumble), which strongly affects both combustion and heat transfer. The op‐ timum intensity of the swirl/tumble rates depends on the type of combustion system, as well as on the specific project targets. As an example, the swirl ratio in DI engine with a bowl in the piston should be high enough to promote the diffusion of the fuel vapor in the chamber. However, an excessive mean turbulence is detrimental to spray penetration, and heat losses increase.

The energy spent to pump the fresh charge across the cylinder is a fundamental parame‐ ter, even if not the only one, to assess the quality of the scavenging system. In order to find a simplified correlation among the average pressure drop across the cylinder (*Δp*) and the main engine parameters, the gas exchange process in a 2-Stroke engine can be idealized as a steady phenomenon, with the piston fixed at bottom dead center and both inlet and exhaust ports partially open, so that the geometric area of each port corre‐ sponds to the average effective area, calculated over the cycle. As a further simplifica‐ tion, the flow is assumed as uncompressible. According to these hypotheses, the mass flow rate across the cylinder can be expressed as:

$$A\_{\rm eff,av} \sqrt{2\rho \cdot \Delta p} = \frac{\rho \cdot \text{DR} \cdot A\_p \cdot \mathcal{U}\_p}{2} \tag{1}$$

Where ρ is the charge density, *DR* is the Delivery Ratio of the engine (ratio of the delivered fresh charge to the reference mass, calculated as the product of charge density to cylinder displacement), *U<sup>p</sup>* is the mean piston speed. *Aeff,av* is the average effective area of all the ports, that can be expressed as:

$$A\_{eff,av} = \frac{1}{\sqrt{1/A\_T^2 + 1/A\_E^2}}\tag{2}$$

Being *AT* the mean effective area of the transfer ports and *AE* the mean effective area of the exhaust ports.

Combining equation 1 and 2, the following expression for *Δp* is found:

**Figure 1.** Typical configurations of 2-Stroke CI high speed engines

154 Advances in Internal Combustion Engines and Fuel Technologies

The optimization of the scavenging process is one of the most challenging task in the de‐ sign of 2-Stroke engines. In fact, the geometry of the ports-cylinder assembly should be defined in order to guarantee a smooth path of the flow across the engine (low flow

**3. Scavenging systems**

$$\Delta p \propto \rho \cdot \text{DR}^2 \cdot \mathcal{U}\_p^2 \cdot \left(\frac{A\_p}{A\_{\text{eff},av}}\right)^2 \tag{3}$$

The following observations can be made:

**1.** Equation (3), despite the simplifications, is able to yield qualitative information about the engine permeability, i.e. the attitude of the ports system to throttle the flow across the cylinder.

while the elevation angle of the rear transfer port should be higher than that of the oth‐ er transfers, to prevent short circuiting. A design optimized for a SI racing engine is shown in figure 2 (right) [16]. For a Diesel engine, this design represents the limit at which to tend for achieving the maximum cylinder permeability. However, since mean piston speed is generally low, it is convenient to reduce the width of the ports (less con‐ cern for piston rings and liner durability) and avoid the overlapping between transfer and exhaust (less risk of short-circuit). When permeability is not an issue at all, a further simplification that can be done is to design just one exhaust port. The advantage is the removal of a quite critical region, from the thermal point of view, i.e. the bridge between

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**Figure 2.** left) sketch of the ports in loop scavenged configuration 1 and (right) ports development of a 125 cc 2-S SI

In another loop configuration, represented in figure 3, the intake system is made up of 2 symmetric manifolds, wrapped around the cylinder, and 2 symmetric sets of 4 inlet ports. This solution is specifically designed for external scavenging. The manifolds cross section width is smaller than the height, in order to reduce the cylinders inter-axle. Furthermore, the cross section area is decreasing along the manifold axis, in order to have a more uniform dis‐ tribution of the flow rate through the inlet ports. It is observed that all the inlet ports are oriented toward one focal point within the cylinder, at the opposite side of the exhaust ports, as suggested also by Blair [17]. The ports are attached to the manifold through short ducts, which have the task of driving the flow towards the cylinder head, for minimizing

In the CFD studies reported in [8] and [14], the most important design parameter for the in‐ let system was found to be the upsweep angle of the ports, see figure 4. As this angle in‐ creases, scavenging efficiency improves, but the port effective area is reduced. The best trade-off depends on a number of specific design issues, so that no general rule can be giv‐ en. In the project described in [8], where the unit displacement of the engine was 350 cc (bore 70 mm, stroke 91 mm, maximum engine speed 4500 rpm), the best results have been

the two exhaust ports.

T1 T2

T3 T4

E1

E2

racing engine by Honda, bore x stroke: 54 x 54.5 mm, EPO/TPO: 82.0/111.6 atdc, [16]

short-circuiting. These ducts have the shape sketched in figure 4.

obtained with an angle of 45° for all the ports.

T5


While permeability is related to the mean piston speed, Diesel combustion is affected by en‐ gine speed: the lower is the maximum number of revolutions per minute, the less is the need of turbulence to support air-fuel mixing.

A number of different lay-outs has been proposed in more than one century of history, and it would be quite hard to review all of them. The two most widespread designs, at least for high speed engines, are the Loop and the Uniflow configurations, the former with piston controlled ports, the latter with exhaust poppet valves, driven by a camshaft, and piston controlled inlet ports. Uniflow scavenging with opposed pistons is not considered, for the sake of brevity.

CFD simulation is the key for the design of modern scavenging systems. The numerical analyses are carried out by means of 3D tools, which are able to predict the flow field details within the cylinder and through the ports under actual engine operating condi‐ tions. Because of the computational cost, the simulation domain is limited to a single cyl‐ inder, and to the portion of cycle included between exhaust port opening and exhaust port closing. Therefore, initial and boundary conditions must be provided by another type of CFD tool, able to analyze the full engine cycle and the influence of the whole en‐ gine lay-out, even if in a simplified manner (in particular, the spatial distribution of the flow through the intake and exhaust systems is considered as one or zero dimensional). The authors have applied this methodology in a number of studies [8, 12-14, 20-25], com‐ paring the simulation results to the experiments, whenever possible. CFD simulation was found to be a quite reliable tool, provided that the numerical models are always under‐ pinned by some experimental evidence.
