Mitsubishi Heavy Industries Technical Review Vol. 52 No. 1 (March 2015)
64

Development of High-efficiency Gas Enginewith Two-stage Turbocharging System

YUTA FURUKAWA*1MINORU ICHIHARA*2

KAZUO OGURA*2

AKIHIRO YUKI*3

KAZURO HOTTA*4DAISUKE TAKEMOTO*4

A new G16NB gas engine for generator use adopting a two-stage turbocharging system wasdeveloped for the improvement of electrical efficiency and output. Two-stage turbocharging, intake
valve close timing and expansion ratio optimization, as well as an increased peak firing pressure,
were adopted to achieve the development target of 10% (4% points) or higher efficiency
improvement. The test engine operated by city gas (13A) displayed the results below: (1)Turbocharger efficiency was improved by 14% points compared with that of conventional models.
In addition, the pressure ratio of compressor was increased by 1.5 contrasted with that ofconventional models at the rated output. (2) The electrical efficiency of the engine was increased
by 4% points, and attained the world’s top level efficiency of 44.7% or higher. Additionally, thebreak mean effective pressure (BMEP) was raised by 0.5 MPa in contrast to that of the current
model.

|1. Introduction
Recently, gas engines for generator use are required to have both high efficiency and highpower. The Miller cycle is well known as one of the means to achieve high efficiency while
avoiding abnormal combustion conditions such as knocking. However, the Miller cycle inherently
lowers the volumetric efficiency, and the output power is lowered at the same boost pressure. A
turbocharger is required to ensure a higher pressure ratio and higher efficiency to compensate forthe disadvantages of the Miller cycle.
Contrarily, a turbocharger tends to lower the turbocharging efficiency in proportion to therise of the pressure ratio at a certain pressure ratio and above, and so a high pressure ratio and highturbocharging efficiency are a trade-off. One solution to achieve both a high pressure ratio and highturbocharger efficiency is two-stage turbocharging with intermediate cooling. In a two-stage
turbocharging system, each pressure ratio at the lower and higher stages can be designed to berelatively low, and the high pressure ratio can be attained at the high-efficiency operational point ofeach stage. Moreover, the effect of intermediate cooling improves turbocharger efficiency.
In two-stage turbocharging, the pressure ratio of each stage can be set to be relatively low, as
can the compression temperature after the compressor becomes low compared with that ofsingle-stage turbocharger. This is advantageous in designing the compressor strength.
This report describes the development of the new G16NB high-efficiency gas engine
adopting a two-stage turbocharging system. The development target of this engine is a higher
efficiency of 10% (4% points) or higher than that of conventional models. In this development,
two-stage turbocharging, the optimization of the inlet valve closing timing (IVC) and expansion
ratio, and a higher peak firing pressure (higher Pmax) are adopted to achieve the desired results.
The development of the two-stage turbocharging system and the design of the component parts ofthe combustion chamber are detailed in this report.
*1 Engine Engineering Department, Engine Division, Machinery, Equipment & Infrastructure*2 Chief Staff Manager, Engine Engineering Department, Engine Division, Machinery, Equipment & Infrastructure*3 Chief Staff Manager, Yokohama Resarch & Development Center, Technology & Innovation Headquarters*4 Yokohama Resarch & Development Center, Technology & Innovation Headquarters

Mitsubishi Heavy Industries Technical Review Vol. 52 No. 1 (March 2015)65|2. Development of G16NB gas engine design
The main specifications of the G16NB engine are shown in Table 1. The output is increased
from 1500kW to 2000kW with the same cylinder bore, stroke length, rated speed and cylinder
numbers. The break mean effective pressure (BMEP), which refers to the load per cycle and unit
displacement volume, is raised by 0.5MPa.
With these specifications, one dimensional performance simulation was conducted using theIVC, target turbocharger efficiency, the target compressor pressure ratio and the target peak firingpressure as parameters to clarify the conditions to achieve the target of efficiency improvement.
GT-Power V7.2 calculation software (Gamma Technologies Inc.) was used for thecalculation investigation. Single-stage turbocharging with a turbocharger efficiency equivalent totwo-stage turbocharging was calculated, although the actual engine uses two-stage turbocharging.
In the calculation, the engine output, expansion ratio and excess-air ratio were set to be constant,and the ignition timing was adjusted to the knocking limit. BMEP was set to the increase target
value by 0.5 MPa. The heat-receiving rate was estimated by extrapolation by conducting asensitivity analysis of BMEP from the test results of the current model.
Table 1
Cylinder bore
StrokeRated engine speed
Cylinder number
Rated output (at generator end)
Break mean effective pressure
Combustion type
Ignition system

2.1

G16NB main specificationsmmmm
min-1－kW
MPa－－Conventional engine Developmental engine
GS16R2
G16NB
170
170
220
220
1500
1500
16161500
2000
1.6
2.1
Pre-chamberPre-chamber
Spark ignition
Spark ignition

Determination of inlet valve close timing, target turbocharger efficiency andtarget peak firing pressure

The calculation results of thermal efficiency, compressor pressure ratio and peak firingpressure (Pmax) to IVC are shown in Figure 1.

Figure 1

One-dimensional performance simulation

As the thermal efficiency calculation shown in the left upper position of Figure 1, theelectrical efficiency improvement cannot be attained by advancing IVC in the case of current
turbocharger efficiency (◆).
The advancement of IVC by 20 deg, the improvement of target turbocharger efficiency byMitsubishi Heavy Industries Technical Review Vol. 52 No. 1 (March 2015)
66

10% points from that of the current turbocharger, the increase of the compressor pressure ratio by
1.5 from the current model, and the increase of Pmax by 5MPa from that of the current model arerequired to achieve the target electrical efficiency.

2.2

Design of two-stage turbocharging system

The two-stage turbocharging system was selected to attain the target turbocharger efficiency
and required pressure ratio clarified in the previous section, and the specifications wereinvestigated.
2.2.1 Determination of parameters
In the target 10% points increase of turbocharger efficiency, 5% points were allotted tothe improvement of single-stage turbocharger efficiency, and the remaining 5% points to thetwo-stage turbocharging system. An outline drawing of the two-stage turbocharging system isshown in Figure 2.
The parameters below are required to design the two-stage turbocharger system.
・Loading ratio of high-pressure compressor
・Intermediate cooling temperature
・Allowable pressure loss of intermediate cooling
Figure 3 shows the results of the parameter study to determine the items above.
The loading ratio of the high-pressure compressor was defined as the required loading
ratio of the high-pressure compressor in the total required compressor loading. In thecalculation, the total compressor pressure ratio including the low and high pressure stages wasconstant, and the single-stage compressor efficiency, turbine efficiency and mechanicalefficiency were constant regardless of the condition.

Figure 2

Outline drawing of two-stage turbocharging system

Figure 3 Parameter case study of two-stage turbocharger system

Mitsubishi Heavy Industries Technical Review Vol. 52 No. 1 (March 2015)
67

Figure 3 shows the results below.
・The optimum point of the loading ratio of the high pressure compressor differs
depending on the intermediate cooling temperature. The approximate value of 0.5
shows the maximum turbocharger efficiency.
・Turbocharger efficiency change is almost linear to the intermediate cooling temperature,and is approximately 0.1% point/K.
・Turbocharger efficiency change is almost linear to the intermediate cooling pressure
loss, and is approximately 0.2% point/KPa.
From the results above, the verification calculation was carried out with the design values
below, and the prospect of satisfying the target turbocharger efficiency was achieved.
・High pressure compressor loading ratio: Approximately 0.5
・Intermediate cooling temperature: Lower than the baseline by 30°C or less・Intermediate cooling pressure loss: Less than 10kPa
2.2.2 Determination of parameters
The ratio of the equivalent area was determined to satisfy the loading ratio of the highpressure compressor investigated in the previous section.
In addition, the high-pressure compressor pressure ratio and volumetric flow rate werecalculated, and the high- and low-pressure stage compressors were designed. The compressor
was designed to have the maximum efficiency near the operating pressure ratio when
considering the low pressure ratio of each single stage. The turbocharger efficiency of a single
stage was expected to rise by 5 % points contrasted with the current model averaging the lowand high-pressure stage compressors.
A design study of the intermediate cooler layout was conducted to satisfy the target
temperatures of the high- and low-pressure stage turbochargers and the intermediate cooler.
Figure 4 shows the results of the pressure loss calculation.
The intermediate cooler pressure loss calculation was made in several layouts, and finallya layout which could satisfy the target value of less than 10kPa was determined.

Figure 4 Pressure loss calculation

2.3

Design of combustion chamber

The flow analysis and flame propagation simulation in the pre-chamber was conducted tooptimize the pre- and main combustion chambers. In addition, thermal structure analysis wascarried out to design a combustion chamber that can be realized considering the mechanicalstrength under the high Pmax and BMEP conditions.
2.3.1 Flow analysis and flame propagation simulation in pre-chamber
The flow analysis in the pre-chamber was conducted to stabilize the combustion in thepre-chamber. Figure 5 shows the methane concentration distribution at the timings of 180 deg
and 90 deg before compression top dead center, and the ignition timing. Figure 5 shows thedilution of the supplied fuel gas by lean mixture flowing from the main chamber, and a uniform
mixture formed at the ignition timing. The optimization of the pre-chamber shape wasconducted using the gas trap rate (fuel gas amount remaining in the pre-chamber at the timingof ignition to the supplied fuel gas) and the flow speed around the ignition plug as theevaluation indicator.
The flame propagation simulation was carried out using the three-dimensional
combustion analysis code for the design of the main chamber shape, and the heat-receiving rate,
the gas temperature in the combustion chamber and the heat transfer coefficient wereMitsubishi Heavy Industries Technical Review Vol. 52 No. 1 (March 2015)
68calculated. For the three-dimensional combustion analysis code, KIVA was used to solve 42
chemical species and 213 elementary reactions.

Figure 5

Flow analysis in the pre-chamber

The calculated heat-receiving rate and isothermal surface in the analysis are shown in
Figure 6. The figure shows the isothermal surface of 1500°C from 5 deg before to 25 deg after
the compression top dead center. The torch flame discharged from the pre-chamber and theflame surface propagation from the torch are shown.
The heat-receiving rate calculated in the analysis was fed back to the one dimensional
performance simulation in section 2.1 to improve the accuracy of performance calculation, andthe gas temperature in the combustion chamber and the heat transfer coefficient calculated in
the three-dimensional combustion analysis were used as the thermal boundary conditions in thestrength analysis of the piston, cylinder liner, and cylinder head.

Figure 6

2.3.2

Flame propagation simulation

Thermal structure analysis of combustion chamber component parts
For the strength analysis of the combustion chamber component parts, the thermalboundary conditions in the combustion chamber were determined by the combustion analysis in
the previous section, and the thermal boundary conditions on the cooling face were determinedby the one dimensional flow network analysis. The MHI flow network analysis code was used
for the flow network analysis of cooling system.
The flow channel cross section, flow channel circumferential length, and flow channel
length were used as parameters, and the cooling water channel of the whole engine wascalculated as a one dimensional model for the flow network analysis.
The flow network analysis can calculate the heat transfer coefficient on the main parts
such as the cylinder liner and cylinder head, in addition to the flow speed calculation at eachchannel. As an example, the developed view of the heat transfer coefficient distribution aroundthe cylinder liner and one dimensional model are shown in Figure 7.
The cooling water passage was designed to equalize the heat transfer coefficient in thecircumferential direction at the upper part of the cylinder liner using the results of the analysis.
FEM analysis was conducted using the determined thermal boundary conditions on thecombustion chamber side and cooling water side, and the cylinder was designed to havesufficient fatigue strength.

Mitsubishi Heavy Industries Technical Review Vol. 52 No. 1 (March 2015)
69Figure 7

Flow network analysis of cooling system|3. Test results
A test engine equipped with a two-stage turbocharging system was built, and a performance
test was conducted. The main measurement results are shown in Table 2 and Figure 8. city gas(13A) was used for the fuel.
The results were sufficient to attain the target values of intermediate cooling temperature andpressure loss in the two-stage turbocharging system. Then, the target values were satisfied by
improving the turbocharger efficiency by 14% points and the compressor pressure ratio by 1.5 incontrast to those of the current model at the rated operational point.
In addition, the relationship of the compressor pressure ratio and turbocharger efficiency atengine partial load was compared. As a result, it was found that the boost pressure rate of increaseat engine starting is advantageous, and the startability is free from problems with the improvementof turbocharger efficiency from the low pressure ratio condition that is used during engine starting.
The electrical efficiency and BMEP of the G16NB engine were increased by 4% points and0.5 MPa, respectively, compared with conventional models with a single stage turbocharger of thesame cylinder bore, stroke length, and engine speed. As a result, the world’s top level electricalefficiency of 44.7% or higher was achieved.
Table 2 Test results
Intermediate cooling temperature
Pressure loss of intermediate cooling⊿Pressure ratio of compressor⊿Turbocharger efficiency⊿Break mean effective pressure⊿Generation efficiency

Figure 8℃kPa－% points
MPa% pointsTarget＜Baseline-30＜10Current model＋1.5＋10＋0.5＋4Measured
Baseline－355
Current model＋1.5＋14＋0.5＋4

Test results|4. Conclusion
The new G16NB generator gas engine was designed, and the test engine displayed the test
results below:(1) The optimization of the two-stage turbocharging system was conducted. Turbocharger
efficiency was raised by 14% points, and the pressure ratio was raised by 1.5 contrasted withthose of conventional engine models.

Mitsubishi Heavy Industries Technical Review Vol. 52 No. 1 (March 2015)
70(2)A two-stage turbocharging system was installed on the engine, and the engine electricalefficiency of the world’s top level of 44.7% or higher was attained with an efficiency
increase of 4% points. Furthermore, the BMEP was raised by 0.5 MPa in contrast to that ofconventional models using city gas (13A) fuel.
This development was conducted with the aid of the fiscal 2012 “Program for Strategic
Innovative Energy Saving Technology” of the New Energy and Industrial Technology
Development Organization (NEDO).
Endurance testing using the verification facilities of our company shown in Figure 9 will be
conducted. In addition, plans call for the engine to be launched on the market after evaluating
engine reliability including the two-stage turbocharging system.

Figure 9 Verification facilities