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A numerical study on the effects of bowl and nozzle geometry on performances of an engine fueled with diesel or bio-diesel fuels

  • Xianyin Leng EMAIL logo , Haiqi Huang , Zhixia He , Qian Wang , Wuqiang Long and Dongsheng Dong
Published/Copyright: July 14, 2022
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Green Processing and Synthesis
From the journal Green Processing and Synthesis Volume 11 Issue 1

Abstract

One of the most important trends for advanced diesel engines is downsizing, i.e., higher power density, which means more fuel is burned in a shorter period. In order to achieve rapid combustion for a high-power density diesel engine, the effects of bowl shape, diameter-to-depth ratio of bowl, and arrangement of nozzle holes on combustion and performance were investigated by CFD simulation. The effects of four bowl shapes, two of which were double-layer split bowls (DLSBs), as well as four diameter-to-depth ratios and three arrangements of nozzle holes were numerically assessed. The results show that the DLSB with a shallow dish-like structure yielded a remarkable effect of swirling flow by fuel splitting into upper- and lower-layer zones, which improved fuel–air mixing, shortened combustion duration, thus, resulting in high combustion efficiency and power density. Moreover, with the increase in diameter-to-depth ratio of the B type DLSB, the turbulent kinetic energy and the peak of pressure and heat release rate increased, further increasing power density. Finally, when the DLSB with a diameter-to-depth ratio of 2.0 is coupled with the staggered double-layer arrangement of nozzle holes, the in-cylinder mixtures became more uniform at both the circumferential and radial directions, and the combustion was considerably accelerated, achieving an optimum specific power of 122.6 kW·L−1. Meanwhile, there was a slight decrement for peak pressure and NOx emission, and smoke decreased by 49.1%, which revealed substantial improvement in reduction in mechanical load and emissions.

Keywords: diesel engine; high power density; piston bowl geometry; double-layer arrangement of nozzle holes; simulation

1 Introduction

Diesel engines are widely used in transportation sector and stationary applications due to their excellent energy efficiency, durability, and energy density [1]. However, as one of the hugest consumers of petroleum products and thereby a major contributor to pollutant particles in the air, diesel engines are facing great challenges relating to global warming and environmental protection [2,3]. To meet these challenges, some researchers suggest to use low-carbon fuels or carbon-free fuels [4], such as natural gas [5,6], methanol [7,8], hydrogen [9], and ammonia [10]. Besides those ways, further improvement of engines fueled with diesel or biodiesel fuel is still a significant way. Biodiesel has similar physical properties with diesel fuel and can be used in diesel engines without much modifications [11]. The use of biodiesel fuel in diesel engines has a lot of advantages, including better lubricity, lower greenhouse gas emissions, and lower pollutant emissions [12,13]. One of the solutions to improve engine performance is to increase power density [14,15], which means a downsized engine for an identical power requirement. And a downsized engine reduces the vehicle’s weight and increases the load capacity, both of which are beneficial for reducing the carbon emissions, making the use of diesel engines or engines fueled with diesel-like e-fuels a greener process.

There are a lot of achievements in the development of high-power density (HPD) diesel engines. The MTU MT 890 series heavy-duty diesel engines reached power density of 92 kW·L−1 [16]. And the BMW company and other manufactures have developed diesel engine with power density over 100 kW·L−1 [17,18]. Diesel engines with even higher power density are in development. With the development of key technologies such as high-pressure common rail system and turbocharger, the speed, brake mean effective pressure, and power density of diesel engines have also been increased. However, there are still many challenges. With the increase in engine speed, diesel engines need to finish combustion in a shorter combustion duration, which requires rapid fuel–air mixing and rapid combustion for HPD diesel engines [19].

The geometry of a piston bowl has a vital impact on fuel–air mixing and combustion, as a good combustion chamber guides the in-cylinder gas flow and enhances air entrainment to sprays improving the performances of diesel engine [20]. On the other hand, in order to improve atomization and spray characteristics, the injection pressure has been largely increased, and the diameter of nozzle holes has been substantially decreased. Nishida et al. [21] investigated the combined effects of increasing the injection pressure to 300 MPa and downsizing the nozzle hole diameter to 50 µm on air entrainment, liquid length, lift-off length, and soot formation of diesel spray flames, and found that the combination use of micro-holes and ultra-high injection pressure greatly accelerated the mixing of fuel and ambient gas, and drastically decreased soot emission. However, when using very fine nozzle holes, to maintain an identical mass flow, the number of holes for a nozzle will increase, leading to interference of neighboring sprays, which in turn results in local rich zone and increased soot emission. In this case, a double layer arrangement of nozzle holes can be helpful. Thus, it is of great significance to study the influence of piston bowl geometry and nozzles arrangement on the flow movement, fuel–air mixtures, and combustion process in the cylinder in order to achieve rapid fuel–air mixing and combustion for HPD diesel engines.

High speed, high boost, and high fuel injection pressure are the main characteristics of HPD diesel engines. Usually, high injection pressure is used in HPD diesel engines to improve fuel atomization and promote the fuel–air mixing, to realize rapid combustion. However, the ultra-high fuel injection pressure also brings some problems, such as spray impingement and fuel accumulation on combustion chamber walls [22]. On the other hand, regarding combustion chamber geometries, the conventional ω-shaped bowl or re-entrant piston bowl with small diameter-to-depth ratio can provide strong squish flow to promote fuel–air mixing, but there are still fuel accumulation at impinging points and poor utilization of the air in the clearance region. Most research works on the bowl geometry of HPD diesel engine focused on the conventional re-entrant piston bowl or ω-shaped bowl, investigating the effects of factors such as the cavity throat diameter and diameter-to-depth ratio [23,24]. Compared to the re-entrant piston bowl or ω-shaped bowl, the double-layer split bowl (DLSB) (Figure 1) proposed by Long et al. [25] can guide the sprays to form diverging flow, significantly suppressing fuel accumulation at the impinging point, leading to rapid dispersion and wide spreading of spray, thereby improving the utilization of air in the clearance region. The effects of the DLSB chamber on a heavy-duty high-speed diesel engine have been verified by several preliminary works. Thanks to the improved fuel–air mixing, the DLSB chamber acquired a 4% reduction in brake specific fuel consumption [26,27]. In view of the advantages of DLSB chamber on improving fuel–air mixing, it is expected that this type of bowl geometry can be used in HPD diesel engines.

Figure 1 The double-layer split combustion system.
Figure 1

The double-layer split combustion system.

The arrangement of nozzle holes has a direct effect on the in-cylinder fuel–air mixing. Compared with the single-layer arrangement of nozzle holes, the double-layer arrangement of nozzle holes contributes significantly in making the spray more uniform in both the axial and radial directions of the combustion chamber [28], and reduces local rich zones, so as to further accelerate the combustion of diesel engine. Wei et al. [29] focused on the study of the double-ω combustion system matched with double-layer arrangements of nozzle holes, and suggested that the double-layer arrangements improved the distribution of spray and the mixture quality in the cylinder, especially the combination of upper-layer 5 holes and lower-layer 3 holes could enhance the combustion performance and further improve the combustion efficiency. Li et al. [30] also reported that properly using double-layer arrangement of nozzles enlarged the diffusion space of spray and improved fuel–air mixing quality, leading to more complete combustion. Due to its advantages in promoting fuel–air mixing rate, the double-layer arrangement of nozzle hole could be applied in HPD diesel engines, and is expected to achieve higher performances.

In view of the advantages of the double layer split combustion chamber and the double-layer arrangement of nozzle holes, in this study numerical simulation was performed to understand the effects of using these two concepts on a HPD diesel engine, with the aim of increasing the specific power from 108 to 120 kW·L−1. Several bowl shapes, diameter-to-depth ratio and nozzle arrangements were numerically studied, and the results were assessed in terms of specific power and combustion efficiency.

2 Numerical methods

2.1 Physical models

A commercial CFD package CONVERGE 2.3 [31] was employed to simulate the physical and chemical processes in the engine cylinder, including turbulent flow, spray, combustion, heat transfer, and the formation of emissions. Table 1 lists the physical and chemical models used in the calculation. Turbulence is of great significance in the work processes of internal combustion engine, especially the ones with HPD, because in HPD engines, the mode of turbulent flow, including swirl flow, squish flow, and reverse squish flow, play critical roles on the rapid fuel–air mixing and combustion processes. Consequently, to use a rational physical model to describe the turbulence behaviors is of crucial importance. In this study, the RNG k–ε model [32] was employed to describe the effects of turbulence and thereby to close the conservation equations. This model is theoretically rational, numerically robust, and extensively applied in the modeling of flow inside internal combustion engine [33]. The recent work of Perini et al. [34] shows that the RNG k–ε model provides the best accuracy trade-off for both swirl and squish flows, applying well to engine-involved turbulences.

Table 1

Models used in the CFD numerical simulation

Description Model
Equation of state Redlich–Kwong [33]
Turbulence model RNG κ–ε [30]
Spray breakup model KH-RT [34]
Drop evaporation model Frossling [35]
Fuel collision model NTC collision [36]
Spray-wall interaction model O´Rourke wall film [37]
Combustion model SAGE [38]

For fuel injection, the breakup of liquid jet is simulated using the Kelvin–Helmholtz model for primary breakup, and Rayleigh–Taylor model for the secondary breakup. The O’Rourke model was applied to resolve spray-wall interaction and depict the wall film flow, and the Frossling model and the NTC model were adopted for the droplet evaporation and the drop collision, respectively [39].

The SAGE model [40] was used to calculate the detailed chemical reaction in each cell according to the pressure, temperature, and mixture concentration of the cell for every time step. In this study, n-heptane was chosen to represent the diesel. And the chemical kinetic mechanism of n-heptane developed at Chalmers University [42], which consists of 42 species and 168 reactions, was selected to calculate the in-cylinder combustion.

2.2 Geometrical models and initial conditions

The current investigation is based on a small-bore heavy-duty diesel engine. The detailed specifications of the engine are listed in Table 2. During the simulation, a closed-cycle starting from the intake valve closing (IVC) timing to the exhaust valve opening (EVO) timing was modeled. Therefore, the in-cylinder geometrical model was built, as shown in Figure 2a. It can be seen that the details of the combustion chamber, including the four valve recesses on the piston top, were precisely modeled.

Table 2

Specifications of the diesel engine

Items Parameters
Bore 98 mm
Stroke 102 mm
Connecting rod 174 mm
Compression Ratio 11:1
Speed 4,100 rpm
Rated power 108 kW
IVC –105° CA ATDC
EVO 120° CA ATDC
Figure 2 The geometrical model and meshes of the engine. (a) The geometrical model and (b) the grid structure (at TDC).
Figure 2

The geometrical model and meshes of the engine. (a) The geometrical model and (b) the grid structure (at TDC).

For the generation of meshes, an adaptive mesh refining technique was employed, in which a relatively coarse mesh was initially generated and then finer meshes are adaptively generated according to the velocity or temperature gradients. Therefore, meshes at several critical regions, such as the boundaries, and the position of spray and flames were refined, as shown in Figure 2b. In this study, after some mesh independent investigations, the baseline mesh size was set as 4 mm, and a three-level refinement was implemented, corresponding to a minimum mesh size of 0.5 mm. And finally, the cell number was ranged from 300,000 to 500,000 during the calculation. The initial pressure and temperature were 0.35 MPa and 378 K, respectively, which were obtained from a one-dimensional CFD modeling. Table 3 provides the injection parameters.

Table 3

Parameters of the injector

Items Parameters
Injector nozzle 12 mm × Φ0.214 mm
Injection pressure 250 MPa
Injection timing –10° CA ATDC
Injection duration 28° CA
Fuel mass 190 mg

2.3 Simulation schemes

Three series of calculations were carried out. First, two designs of the DLSB bowl (DLSB-A and DLSB-B) were compared with the original piston bowl (a re-entrant bowl) and the ω-shaped bowl, as shown in Figure 3a. The differences between DLSB-A and DLSB-B bowls were as follows. The upper layer of the DLSB-B bowl has a concave shape, like a shallow dish, while that of the DLSB-A bowl was a convex shape. Moreover, the curve near the central of the bowl profile was convex and concave for the DLSB-A bowl and DLSB-B bowl, respectively. Based on the numerical results, an optimum bowl shape was selected (here the DLSB-B bowl), and then a thorough investigation of the effects of the diameter-to-depth ratio of this type of chamber bowl was carried out, which is the second series calculation. The different diameter-to-depth ratios of the bowl investigated are shown in Figure 3b. And then, the optimum diameter-to-depth ratio was obtained based on the numerical results. Finally, for the optimum bowl geometry, the effects of nozzle hole arrangement were numerically assessed. Figure 4 illustrates the three arrangements of nozzle holes, namely, uniform single-layer, overlapped double-layer, and staggered double-layer. The details of the geometrical parameters of the bowls in this study are listed in Table 4.

Figure 3 Bowl geometries. (a) Different bowl shapes and (b) different diameter-depth ratios of DLSB.
Figure 3

Bowl geometries. (a) Different bowl shapes and (b) different diameter-depth ratios of DLSB.

Figure 4 Different arrangement schemes of three nozzle holes based on DLSB-B.
Figure 4

Different arrangement schemes of three nozzle holes based on DLSB-B.

Table 4

Parameters of simulation schemes

Schemes Parameters
Piston bowls Re-entrant, ω-shaped, DLSB-A, DLSB-B
Diameter-to-depth ratios D/H B0:1.4 B1:1.6 B2:1.8 B3:2.0
Injection angles 152° 148° 144° 140°
Arrangement of nozzle holes Uniform single-layer, overlapped double-layer, staggered double-layer
Injection angles of upper and lower layers were, respectively, 160° and 100°

In order to avoid the interference of the upper-layer and lower-layer fuel sprays of the double-layer arrangements of nozzle holes at the axial directions, the including angle of lower-layer holes were set at a smaller value, as shown in Figure 4. Considering that there were more cavitations in the upper-layer holes than the lower-layer holes [43], the fuel injection quantities of upper-layer holes were set to be 4% less than that of the lower-layer holes, which is equivalent to the discharge coefficients of upper-layer nozzle that was about 3.4% lower than that of the lower-layer nozzle in this work. Furthermore, under the limit of maximum pressure of 21 MPa, the injection timing of all calculating schemes needs to be fixed at −10° CA ATDC.

The specific power in unit of kilowatt per liter is one of the most significant performance indicators of HPD diesel engines. And the uniformity of fuel–air mixtures is of crucial importance for the rapid combustion of HPD diesel engines. Therefore, in this study, the specific power was extracted from the calculated results and the uniformity of the mixture (the standard deviation [STD] of the equivalence ratio) was quantified for comparison. These parameters are given by:

(1) P L = η m P i V s

(2) ∅ STD = ∑ cell m cell ( ∅ cell − ∅ mean ) ∅ total ,

where P L, P i, η m, and V s represent specific power, indicated work, mechanical efficiency, and displacement volume, respectively. P i was calculated from the integration of the in-cylinder pressure over the in-cylinder volume, and η m adopted a value provided by the manufacture. And the subscript cell indicates cell value, mean represents mean value, total indicates total value, the parameter Φ indicates equivalence ratio and m is the mass.

2.4 Validation of models

2.4.1 Image-based validation

The combustion of HPD diesel engines highly depends on fuel–air mixing processes. Therefore, the precise modeling of spray was the basis of in-cylinder combustion simulation. To verify the credibility of the spray models, the diesel spray processes in a constant volume vessel were calculated, and the numerical results and experimental results were compared in terms of liquid penetration and spray cone angles. Table 5 lists the conditions of the high speed photography experiments of the spray, which were conducted by Xiang et al. [43].

Table 5

Conditions of spray validation

Project Parameter
Holes diameter 0.14 mm
Injection duration 3.6 ms
Common rail pressure 140 MPa
Ambient gas N2
Ambient temperature 750 K
Ambient pressure 2.7 MPa

Figure 5 illustrates the comparison of experimental spray image and the numerically obtained spray image at 2.0 ms after the start of injection (ASOI) under high ambient temperature (750 K). A good agreement of the liquid spray morphology can be observed.

Figure 5 Comparison of experimental and numerical results of the spray (at 2.0 ms ASOI).
Figure 5

Comparison of experimental and numerical results of the spray (at 2.0 ms ASOI).

Figure 6 shows the temporal evolution of the experimental and numerical results of the liquid phase penetration and cone angle of the spray. It is clear that the numerical results reasonably agreed with the experimental results, in terms of both spray penetration and spray cone angle, indicating a good credibility of the spray models.

Figure 6 Spray penetrations and spray cone angles from experiment and simulation.
Figure 6

Spray penetrations and spray cone angles from experiment and simulation.

2.4.2 In-cylinder combustion

To validate the turbulence and combustion model, the in-cylinder flow, spray, mixing, and chemical reaction processes in a 110 mm bored HPD diesel engine were calculated. The operation speed, injection pressure, and boost ratio were 3,800 rmp, 180 MPa, and 4, respectively. Figure 7 shows the comparison of in-cylinder pressure and heat release rates between the experimental and numerical results. It is found that the relative errors between the calculated and experimental in-cylinder pressures were within 1.0%. And it can be seen from the profiles of heat release rates that the starting phase of combustion, the peak heat release rate, and the combustion duration were also well predicted. The predicting precises of both the spray and combustion characteristics were state of the art, similar to those in the studies by Ener et al. [44] and Wang et al. [45]. Therefore, it can be concluded that the selection of numerical models as well as the setting of model parameters were reasonable for the study of HPD diesel engines.

Figure 7 Comparison of experimental and numerical results of in-cylinder pressure traces and heat release rates.
Figure 7

Comparison of experimental and numerical results of in-cylinder pressure traces and heat release rates.

3 Results and discussion

3.1 Effects of piston bowl shape

To understand the effects of piston bowl shape, the in-cylinder combustion processes of four shapes of piston bowl were calculated. Figures 8 and 9 depict the effects of the piston bowl shapes on in-cylinder pressure, heat release rates, and combustion duration. The peak of pressure and heat release rate of the ω-shaped bowl were the highest, followed by the DLSB-B bowl, which reflected the burning rate of the former half stage of the combustion process. In fact, it can be observed at the initial stage of combustion, there were very limited differences in the heat release rates of the four bowl shapes, as revealed in Figure 9, the durations of CA0-10 and CA10-50 were almost the same. It should be due to that at the initial stage, most combustion tools placed inside the bowl space, were marginally influenced by the bowl shape. After 20° CA ATDC, both DLSB-A and DLSB-B bowls showed evident advantages over the other two bowls in enhancing the combustion rate, indicating an acceleration of fuel–air mixing. It is found that the latter half stage of the combustion period (CA50-90) of ω-shaped bowl was the longest one, which should be owing to the fuel diffusion to the periphery of the combustion chamber was inhibited by the narrow top clearance and the air utilization level at the top clearance was poor.

Figure 8 In-cylinder pressure and heat release rate traces under different piston bowls.
Figure 8

In-cylinder pressure and heat release rate traces under different piston bowls.

Figure 9 Combustion durations for different piston bowls.
Figure 9

Combustion durations for different piston bowls.

Figure 10 shows the influence of the piston bowl shape on the specific power and combustion efficiency. It can be seen that the re-entrant bowl yielded the lowest specific power and combustion efficiency, indicating that this bowl structure was not proper for the HPD diesel engine. Compared with the re-entrant piston bowl, the specific power and combustion efficiency of the other piston bowls were substantially improved. The specific power increased to 116 kW·L−1, and the combustion efficiency increased to higher than 90%. The combustion and performance of the DLSB-B bowl were evidently higher than those of the DLSB-A bowl. Regarding combustion efficiency, both the ω-shaped bowl and the DLSB-B bowl obtained a combustion efficiency of around 96%. However, the specific power of the ω-shaped bowl was slightly lower, which should be due to its relatively long combustion duration. To sum up, it seems that the DLSB-B bowl yielded the optimum performance. The next paragraphs will explain these phenomena through detailed information of the in-cylinder processes.

Figure 10 Specific power and combustion efficiency for different piston bowls.
Figure 10

Specific power and combustion efficiency for different piston bowls.

Figure 11 presents the temporal evolution of in-cylinder distribution of equivalence ratio and streamlines, which reveals the integration of fuel–air mixing and gas flow. It can be observed that for the re-entrant piston bowl, the piston wall did not split the sprays effectively, leading to most rich mixture accumulated on the upper zone of the bowl. While for the ω-shaped bowl, there was a narrow top clearance near top dead center, which inhibits the spreading of fuel–air mixture entering the top zone, leading to extended late combustion duration. It is also observed that for the ω-shaped bowl weak swirling flows were formed in the area of the valve recesses, which locally expanded the effective thickness of the top clearance so that the maximum pressure was slightly higher during the former half stage of the combustion period (CA10-50). In both of the double-layer split chambers, the fuel spray was split into upper and lower streams by the fuel impinging platform of the bowls. Furthermore, there were obvious swirling flows in the upper and lower zones under the effect of reverse squish flow, which was beneficial for air entrainment.

Figure 11 The distribution of equivalence ratios and streams at 5° CA ATDC and 10° CA ATDC.
Figure 11

The distribution of equivalence ratios and streams at 5° CA ATDC and 10° CA ATDC.

At 10° CA ATDC, while the fuel accumulated on the bowl wall in ω-shaped bowl, the DLSBs could quickly split and diffuse the fuel through the collision platform, which is also consistent with the observation of Qi and Long [41]. Compared with the DLSB-A bowl, a shallow dish-like structure was applied to the upper and lower geometry of the DLSB-B bowl. In contrast to the DLSB-A bowl, the upper fuel in DLSB-B bowl was not easy to spread along the surface of the upper piston geometry, thereby improving the fuel–air mixing with reverse squish flow. The shallow dish-like structure of the lower piston geometry was helpful to rapid fuel dispersion since the ultrahigh speed leads to insufficient mixing time. Overall, mixing and combustion were further improved by the combination of the DLSB and the shallow dish-like structure in both the upper and the lower zones of the DLSB-B bowl.

Figure 12 illustrates the in-cylinder mean turbulent kinetic energy (TKE) and the STD of the equivalence ratio. Unipolar zone division of the re-entrant piston bowl and ω-shaped bowl gave a good insight as to why the TKE of the DLSB bowls after top dead center was lower than that of the conventional bowls, and the stronger turbulent was caused by the combination of squish flow and direct fuel injection in the zone. A remarkable characteristic of the DLSB bowls was that the combustion chambers were divided into two regions of upper and lower zone. Although the flow disturbance was reduced through fuel splitting at the fuel impinging platform of the bowls, the fuel–air mixing and combustion at the upper zone was substantially improved due to better utilization of the fresh air in the clearance zone. Nevertheless, the superior performances of the DLSB-B bowl indicated that fast fuel dispersion was more important than turbulent disturbance in HPD diesel engines at high speeds. It is well-known that the homogeneity of the fuel–air mixture can be quantitatively assessed by the STD of equivalence ratio. The STD of equivalence ratio of the re-entrant piston bowl remained at high level, indicating a relatively poor uniformity of the mixture in the re-entrant piston bowl. During the period of 0–20° CA ATDC, there was little interaction between fuel and chamber wall, and the STD of equivalence ratio of the ω-shaped bowl was lower than that of the DLSB. However, after the 20° CA ATDC, especially under the diversion of the type B DLSB, the mixing was further improved and thus, the STD of the equivalence ratio of the DLSB-B bowl was reversed and became smaller, which meant relief of afterburn in the ω-shaped bowl. In a word, the STD of the equivalence ratio of the DLDC-B bowl was relatively low in the whole combustion duration and had the advantage of uniform mixtures.

Figure 12 TKE and STD of equivalence ratio under different piston bowls.
Figure 12

TKE and STD of equivalence ratio under different piston bowls.

Figure 13 shows soot and NOx emissions under different piston bowls. Soot emission of the re-entrant piston bowl was the highest (over 2.2 g‧kW−1·h−1) due to the rich mixture, meanwhile its poor combustion characteristic led to lower in-cylinder temperatures, resulting in lower levels of NOx emission. Soot emissions can be greatly reduced (to less than 0.43 g‧kW−1·h−1) by rapid fuel diversion with the use of the DLSBs. However, compared with the DLSB-A bowl, soot emission of the DLSB-B bowl with a shallow dish-like structure was decreased by 0.1 g‧kW−1·h−1 because of a minimal mixing advantage on enhancing fuel flow of the upper and lower zones, but NOx emission increased by 0.4 g‧kW−1·h−1. In addition, it can be seen that soot emission level of the ω-shaped bowl was the same as that of the DLSB-B bowl and NOx emission reduced by 0.2 kW−1·h−1. Actually, mixing characteristic of the ω-shaped bowl was worse than that of the DLSB-B bowl, leading to a slower combustion rate and longer combustion duration. Therefore, it resulted in smaller NOx emissions and also enhanced the oxidation of soot during the afterburn period.

Figure 13 Soot and NOx emissions under different piston bowls.
Figure 13

Soot and NOx emissions under different piston bowls.

In brief, mixing of the DLSB-B bowl was better than that of the ω-shaped bowl, combustion was accelerated to shorten combustion duration and to accelerate afterburn, the specific power was 1.4 kW·L−1 higher. It is obvious that the main objective of this study was to achieve fast combustion in a high-speed diesel engine to improve power density. Although the NOx emissions were slightly higher, the DLSB-B bowl could still be selected for further optimization.

3.2 Effects of diameter-to-depth ratio

Figure 14 provides the traces of in-cylinder pressure and heat release rates for the four different diameter-to-depth ratios (ranging from 1.4 to 2.0) of DLSB-B bowls. It is clear that the overall trend was with the increase in the diameter-to-depth ratio, and there was a considerable increase in the combustion rate during the period 0–20° CA ATDC, leading to higher in-cylinder pressure. As mentioned in the previous section, from 0° CA to 20° CA ATDC, there was still little fuel-wall interaction, and most of the fuel–air mixing took place inside the chamber space. Consequently, as the diameter-to-depth ratio increased, the distance from nozzle to the piston wall increased, leading to more air entraining into the spray before the fuel-wall interaction, thereby, causing a higher heat release rate. It is also found that as the diameter-to-depth ratio increased to 2.0, the increase in the combustion rate became marginal, indicating there was no need to further increase the diameter-to-depth ratio.

Figure 14 The traces of in-cylinder pressure and heat release rates for different diameter-to-depth ratios.
Figure 14

The traces of in-cylinder pressure and heat release rates for different diameter-to-depth ratios.

Figure 15 depicts that the combustion duration in terms of CA0-10, CA10-50, and CA50-90 for the four diameter-depth ratios. It can be seen that with the increase in the diameter-to-depth ratio, there was no change in the ignition delay period (CA0-10) and a slight decrease in the former half of the main combustion period (CA10-50), and substantial decrease for the latter half stage of the main combustion period (CA50-90). It is clear that increasing the diameter-to-depth ratio mainly improved the latter half stage of the combustion processes. The traces of the in-cylinder mean TKE and the STD of the equivalence ratio are presented in Figure 16. As the diameter-to-depth ratio increased, the mean TKE increased and the STD of the equivalence ratio decreased throughout the combustion process, implying more uniform mixtures for the cases of higher diameter-to-depth ratios. As it is evident from Figure 16, the increase in turbulent intensity and mixture homogeneity as the diameter-to-depth ratio increased from 1.4 to 2.0 gave a good insight as to why there were higher in-cylinder pressure and heat release rate during 0–20° CA ATDC (Figure 8). After 20° CA ATDC, although the mean TKEs for the cases of all diameter-to-depth ratios decreased, more dispersed mixtures were still seen with increase in the diameter-to-depth ratio due to the mixture-guiding effects of the DLSB bowl, further resulting in shortening of the latter half stage of the combustion period (Figure 9).

Figure 15 Combustion durations for different diameter-to-depth ratios.
Figure 15

Combustion durations for different diameter-to-depth ratios.

Figure 16 The mean TKE and STD of equivalence ratio for different diameter-to-depth ratios.
Figure 16

The mean TKE and STD of equivalence ratio for different diameter-to-depth ratios.

The specific power and combustion efficiency for cases of different diameter-to-depth ratios are illustrated in Figure 17. A substantial increase in the specific power was observed with the increase in the diameter-to-depth ratio. It should be attributed to the acceleration on the latter stage of the combustion, which allowed more energy to be transferred to power. For the combustion efficiency, it was noted that there was evident increase with the diameter-to-depth ratio changing from 1.4 to 1.6. However, with further increase in the diameter-to-depth ratio from 1.6 to 2.0, the combustion efficiency slightly decreased. This phenomenon could be owing to that with the increase in the diameter-to-depth ratio, the height of clearance decreased, therefore, the depth of the valve recesses increased, which influenced the gas flow mode, weakening the splitting effects of the DLSB bowls. Wang et al. [46] also found that with the increase in the depth of the valve recesses from 0 to 8 mm, especially after the depth exceeds 4 mm, the combustion performance deteriorated due to the poor matching of spray and bowl geometry.

Figure 17 Specific power and combustion efficiency under different diameter-to-depth ratios.
Figure 17

Specific power and combustion efficiency under different diameter-to-depth ratios.

Figure 18 shows soot and NOx emissions under different diameter-to-depth ratios. With the increase in the diameter-to-depth ratios, NOx emissions increased and soot emissions decreased, but its decrease was less than 0.01 g‧kW−1·h−1 when the diameter-to-depth ratio exceeded 1.6. Obviously, the increase in the diameter-to-depth ratio strengthened the coordination between the spray and the chamber, further accelerating the fuel diversion and combustion, resulting in the increase in the high temperature region, thus forming more NOx. However, the shape of the combustion chamber was elongated and valve recesses became deeper, the fuel was far away from the center of the combustion chamber after the end of the main combustion period, and the oxidation of soot was slower for diameter-to-depth ratio from 1.6 to 2.0.

Figure 18 Soot and NOx emissions under different diameter-to-depth ratios.
Figure 18

Soot and NOx emissions under different diameter-to-depth ratios.

3.3 Effects of the arrangement of nozzle holes

Figure 19 displays the traces of in-cylinder pressure and heat release rates for the three different arrangements of nozzle holes, namely, uniform single-layer, overlapped double-layer, and staggered double-layer. Compared with the uniform single-layer arrangement of nozzle holes, the overlapped double-layer arrangement of nozzle holes substantially decreased the in-cylinder pressure and heat release rates, which was probably due to the interference of the sprays between layers. For the staggered double-layer arrangement of nozzle holes, the former half of heat release rates was slightly lower than those of the uniform singer-layer arrangement, while the latter half of the heat release rates was evidently higher than those of the uniform singer-layer arrangement. As a result, the in-cylinder peak pressure was slightly lower and higher than that of the uniform single-layer arrangement case at the former and latter halves of the combustion processes, respectively. This indicates that the staggered double-layer arrangement of nozzle holes improved the air–fuel mixing, especially for the later stage of the combustion processes. Similar phenomenon occurred in Bergstrand’s study [43], which demonstrated that a nozzle with double-layer staggered arranged holes enhanced the in-cylinder air utilization, and resulted in higher heat release for both the premixed combustion and the mixing-controlled combustion.

Figure 19 In-cylinder pressure and heat release rate traces under different arrangements of nozzle holes.
Figure 19

In-cylinder pressure and heat release rate traces under different arrangements of nozzle holes.

Figure 20 shows the combustion duration in terms of CA0-10, CA10-50, and CA50-90 for the three arrangements of nozzle holes. Similar to the results of the different bowl shapes mentioned above, the combustion delay periods (CA0-10) of the three arrangements were almost the same. However, the former half stage of the combustion period (CA10-50) and the latter half stage of the combustion period (CA50-90) with the overlapped double-layer hole nozzle were both relatively longer, indicating a decrement in the fuel–air mixing. For the staggered double-layer arrangement of nozzle holes, the latter half stage of the combustion period (CA50-90) was evidently shortened, which was even shorter than that of the uniform single-layer arrangement of nozzle holes. It revealed that for the double-layer staggered arrangement of nozzle holes, the fuel-air spreading in the latter half stage of the combustion period was remarkably enhanced, leading to a rapid combustion rate of that period. This was the advantage of the double-layer staggered arrangement of nozzle holes.

Figure 20 Combustion durations of different arrangements of nozzle holes.
Figure 20

Combustion durations of different arrangements of nozzle holes.

Figure 21 displays the equivalence ratio contours of the three arrangements of nozzle holes. Clearly, at 15° CA ATDC, each pair of sprays emerging from the upper and lower layers of the overlapped double-layer arrangement of nozzle holes showed strong interaction and even merged into a single spray, resulting in local fuel-rich zones near the center of the bowl and decreasing the spray penetrations. This should be exactly the reason for the substantially lower heat release rate at this timing. For the staggered double-layer arrangement of nozzle holes, the lower layer of sprays was directed to the bottom of the chamber bowl, so as to entrain the air in the central region of the bowl. And then, under the guiding effects of the bowl geometry, the tip of the lower layer of sprays moved to the central region of the upper zone, filling the gaps in the upper sprays. On the other hand, the upper layer sprays were split by the bowl and mostly spread to the clearance region, mainly to entrain fresh air from the outer region of the chamber. While at 50° CA ATDC, which was in the latter stage of combustion processes, the in-cylinder mixtures get more uniform for the three arrangements of nozzle holes. However, the in-cylinder distribution characteristics for the three arrangements were largely different. For the case of uniform single-layer arrangement of nozzle holes, the mixtures were mainly concentrated on the periphery zone of the combustion chamber, resulting in poor air utilization of the central zone. And for the case of overlapped double-layer arrangement of nozzle holes, most of the mixtures were concentrated in the central zone of the combustion chamber due to the interaction of double-layer sprays, which made it difficult for the mixture to spread to the periphery zone. While for the cases of staggered double-layer arrangement of nozzle holes, it is found that mixtures were more uniform in both the circumferential and radial directions of the combustion chamber, implying a further improvement in air utilization, which accelerated combustion of the latter stage of combustion processes compared to the single-layer arrangement of nozzle holes. Bergstrand and Denbratt [47] also reported that the staggered double-layer arrangement of nozzle holes could reduce interference and overlap among the adjacent fuel sprays, and improve air utilization and spatial distribution of fuel.

Figure 21 The equivalence ratio contours for different arrangements of nozzle holes.
Figure 21

The equivalence ratio contours for different arrangements of nozzle holes.

Figure 22 shows the specific power and combustion efficiency for the three different arrangements of nozzle holes. Compared with the uniform single-layer arrangement of nozzle holes, the combustion efficiency of the overlapped double-layer arrangement of nozzle holes was reduced due to spray interference and poor utilization of the fresh air in the peripheral zone. On the contrary, the combustion efficiency of the staggered double-layer arrangement of nozzle holes reached 99.3% as both the radial and circumjacent distribution of mixture was improved, especially in the latter stage of combustion processes. As a result of the improved fuel–air mixing, the specific power of the staggered double-layer arrangement of nozzle holes obtained a specific power of 122.6 kW·L−1, which achieved the initial aim of this study.

Figure 22 Specific power and combustion efficiency for different arrangements of nozzle holes.
Figure 22

Specific power and combustion efficiency for different arrangements of nozzle holes.

Figure 23 shows soot and NOx emissions for different arrangements of nozzle holes. The bad mixing and combustion restrained heat release of fuel to form NOx, while the rich mixtures with the spray interference was the main reason of a higher soot emission level under the overlapped double-layer arrangement of nozzle holes. Compared with the uniform single-layer arrangement of nozzle holes, soot and NOx emissions of the staggered double-layer arrangement of nozzle holes decreased by 49.1% and 20.4%, respectively. The reason was that both the radial and circumjacent distribution of mixture were improved to decrease rich mixtures by the staggered double-layer arrangement of nozzle holes, resulting in lower soot emission (0.16 g‧kW−1·h−1). On the other hand, the upper volume of the DLSB was larger due to four valve recesses. Therefore, the reduction in spray density into the upper space with the staggered double-layer arrangement of nozzle holes led to a slight decrement in the heat release rate at the beginning of the combustion, and both the in-cylinder temperature and peak pressure were relatively lower, which encouraged to reduce NOx emission and mechanical load.

Figure 23 Soot and NOx emissions for different arrangements of nozzle holes.
Figure 23

Soot and NOx emissions for different arrangements of nozzle holes.

4 Conclusion

The geometry of double-layer split combustion chamber and double-layer arrangement of nozzle holes were numerically evaluated for further increasing the specific power of the HPD diesel engine. The following conclusions were drawn based on this work.

  1. The design of double-layer split chamber was able to split the sprays into the lower layer and upper layer of the combustion chamber, accelerating spray dispersion and mixing，which hindered afterburn and obtained shorter combustion duration than that of the conventional bowls.

  2. The B-type double-layer split chamber, which was characterized with shallow dish-like structures in both the upper layer and in the center of the bowl, enhanced the fuel behavior and air utilization in both the central zone and peripheral zone of the combustion chamber, yielding shorter combustion duration and higher specific power by 118.7 kW·L−1.

  3. For the B-type double-layer split chamber, with increase in the diameter-to-depth ratio, the spatial mixing increased and near wall mixing decreased, leading to more rapid combustion and higher specific power; however, the combustion efficiency and soot emission marginally decreased due to the deeper valve recesses, and NOx emission increased.

  4. The use of the B-type double-layer split chamber with a diameter-to-depth ratio of 2.0, together with a staggered double layer arrangement of nozzle holes led to more uniform distribution of sprays in both the circumferential and radial directions, thereby, accelerating the combustion rate, achieving an optimum specific power of 122.6 kW·L−1. Moreover, there was a slight decrement in peak pressure and NOx emission, and smoke decreased by 49.1%, which revealed obvious improvement in reduction of structural load and emissions.

  1. Funding information: Jiangsu Provincial Key Research and Development Program, China (BE2019009-5), the National Natural Science Foundation of China (Grant numbers: 51776088, 51779044, and 51876083), and a Project Funded by the Priority Academic Program Development of Jiangsu High Education Institutions.

  2. Author contributions: Xianyin Leng: conceptualization, methodology, funding acquisition, and writing – review and editing; Haiqi Huang: investigation, data curation, and writing – original draft preparation; Zhixia He: methodology, investigation, and visualization; Qian Wang: data curation and funding acquisition; Wuqiang Long: conceptualization, supervision, and funding acquisition; Dongsheng Dong: investigation and experiments.

  3. Conflict of interest: Authors state no conflict of interest.

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Received: 2022-02-02
Revised: 2022-04-08
Accepted: 2022-04-20
Published Online: 2022-07-14

© 2022 Xianyin Leng et al., published by De Gruyter

This work is licensed under the Creative Commons Attribution 4.0 International License.

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  97. Anti-asthmatic activity of Saudi herbal composites from plants Bacopa monnieri and Euphorbia hirta on Guinea pigs
  98. Embedding green synthesized zinc oxide nanoparticles in cotton fabrics and assessment of their antibacterial wound healing and cytotoxic properties: An eco-friendly approach
  99. Synthetic pathway of 2-fluoro-N,N-diphenylbenzamide with opto-electrical properties: NMR, FT-IR, UV-Vis spectroscopic, and DFT computational studies of the first-order nonlinear optical organic single crystal
https://doi.org/10.1515/gps-2022-0054
Keywords for this article
diesel engine; high power density; piston bowl geometry; double-layer arrangement of nozzle holes; simulation
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  97. Anti-asthmatic activity of Saudi herbal composites from plants Bacopa monnieri and Euphorbia hirta on Guinea pigs
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