CC BY
11
----------, :::::
-□ □-
У результатi проведених теоретичних дослiджень виконан розрахунки теоретич-ног тдикаторног питомог витрати пали-ва для рiзних способiв конвертацп на газ дизельних двигутв. Розглядались традицш-ний спо^б дефорсування при конвертацп на газ дизельних двигутв за рахунок встанов-лення додаткових прокладок мiж головкою i блоком цилiндрiв, i запропонований споЫб дефорсування за рахунок затримки закрит-тя впускного клапана.
У результатi проведених в лабораторних умовах експериментальних дослiджень при рiзних способах конвертацп дизельного дви-гуна модифшацп Х170ТЬ легкового автомо-бшя Опель встановлено, що при використанш традицшного способу дефорсування двигутв за рахунок встановлення додаткових прокладок мiж головкою i блоком цилiндрiв значен-ня тдикаторног питомог витрати зросло, в середньому, на 8-9 %. Але при зниженн сту-пеня стиснення двигуна запропонованим способом за рахунок затримки закриття впускного клапана, експериментальне значення тдикаторног питомог витрати не т^ьки не зб^ьшилось, а й зменшилось, в середньому, на 7-8 %. Зниження ступеня стиснен-ня двигуна за рахунок затримки закриття впускного клапана запропонованим способом здшснювалось шляхом змти форми кулачтв розподтьного валу. Для цього кулачки впус-кних клапатв розподтьного валу наплавлялись, а тсля цього шлiфувались до одержання необхiдного проф^ю, при якому вiдбувалась затримка закриття клапатв у потрiбних межах.
Одержан результати дозволяють оптимi-зувати процеси переведення на газ дизельних двигутв та знизити витрату палива конвер-тованих двигутв, в середньому, на 15-17 % у порiвняннi з газовими двигунами, переоблад-нання яких здшснено традицшним шляхом.
Ключовi слова: альтернативн палива, дизельний двигун, конвертация двигуна на газ,
питома витрата палива -□ □-
::::::::::::::: н
UDC 62-614
IDOI: 10.15587/1729-4061.2018.13935811
FUEL ECONOMY RAISING OF ALTERNATIVE FUEL CONVERTED DIESEL ENGINES
S. Kryshtopa
Doctor of Technical Sciences, Associate Professor Department of Automechanical Nadvirna College of the National Transport University Soborna str., 177, Nadvirna, Ukraine, 78400 E-mail: auto.ifntung@ukr.net M. Panchuk PhD, Associate Professor Department of welding structures and restoration of machine parts** E-mail: 24berkyt@rambler.ru F. Ko z a k PhD, Professor* E-mail: auto.ifntung@gmail.com B. Dolishnii PhD, Associate Professor* E-mail: rettes@mail.ru I. Mykytii Postgraduate student* E-mail: 7ivan1@i.ua O. Skalatska PhD
Department of Publishing and Editing Odessa I. I. Mechnikov National University Dvoryanskaya str., 2, Odessa, Ukraine, 65082 E-mail: kr.sv@ukr.net *Department of Oil and Gas Technological Transport** *Ivano-Frankivsk National Technical University of Oil and Gas Karpatska str., 15, Ivano-Frankivsk, Ukraine, 76019
1. Introduction
With current production volumes and proved reserves, oil will be enough for mankind for about fifty years. The second energy resource after oil as motor fuel is natural gas. Recently, there has been a tendency of conversion of existing diesel engines to gas and other alternative fuels, but at the moment a number of issues remain unaddressed by producers [1].
In addition, it is indisputable that cars with engines that operate on gas and other alternative fuels have worse fuel consumption compared to similar diesel engines [2]. At present, the issues of economic feasibility in the conversion to alternative fuels in the present conditions are of paramount importance. Therefore, achieving high fuel economy of gas engines is one of the main requirements in the direction of expanding the use of gas fuels.
In general, for the conversion of diesel engines to gas fuel, it is necessary to do the following:
- to install gas cylinder equipment;
- to mount the ignition system;
- to adjust the gas engine control system and optimize ignition timing angles;
- to reduce the compression ratio for the base diesel engine.
The most common solutions to reduce the compression ratio of the base diesel engine in conditions of automobile plants are:
- the use of reduced-compression-height pistons;
- reduction of the piston stroke by replacing the crankshaft.
Easier ways to reduce the compression ratio of the diesel engine in conditions of service stations or autoenterprises are [3]:
- increase in the combustion volume in the head or pistons by milling;
- increase in the thickness or number of gaskets between the engine block and the head.
In this case, the geometric compression ratio is provided in the range of 12-13 units. The specified reduction in the compression ratio leads to a decrease in the indicated efficiency of the gas engine compared to the base diesel engine. In addition, reduction in the indicated efficiency, especially at low loads, causes an increase in throttling losses. This leads to a deterioration of the gas fuel economy, on average, by 10-25 % compared to the base diesel engine. Therefore, a more up-to-date method of reducing the compression ratio of the base diesel engine due to later intake valve closing is proposed. This allows achieving a reduction in the effective compression ratio, making it considerably smaller than geometric. Therefore, work in this direction is undoubtedly relevant. This allows minimizing the problem of detonation combustion of the gas-air mixture and at the same time increasing the indicated engine efficiency and reducing fuel consumption.
2. Literature review and problem statement
Currently, most of the major car manufacturers consider comprehensive solutions to improve the performance of engines: valve timing variation, intake system improvement, engine compression ratio optimization, etc.
Thus, performance improvement of the MTZ-80 wheeled tractor ( Belarus) during the conversion of its engine to gas fuel has been studied in [4]. The feature of the technology of converting this tractor engine to gas was the fact that the reduction in the compression ratio was due to the installation of additional gaskets between the cylinder head and block. It should be noted that this method of reducing the engine compression ratio leads to an increased risk of blowing of gaskets between the cylinder head and block.
In [5], the conversion of a diesel engine car to run on natural gas has been studied. In the engine, spark plugs, the electronic ignition system and gas cylinder equipment are installed instead of spray nozzles. The combustion chamber in the piston heads was bored and the compression ratio was reduced from 21 to 13 units. However, the authors have not sufficiently considered the processes of heat exchange in the converted diesel engine.
The study of regulation of the admission period in the gas-converted diesel engines has been made in [6]. It has been found that when the intake valve closing is changed, it
is expedient to simultaneously correct the admission period through the valve lift adjustment. But a significant design complication and the associated general significant increase in the cost of the converted engine remained unaddressed.
In [7], the laws of changes in the compression ratio and the intake valve closing and exhaust valve opening angles, depending on the engine load, have been studied. The geometric compression ratio was controlled by a moving-head piston. Comparative calculations of load characteristics of the base and converted engines have been performed. An interesting feature of [7] is the comparison of working processes, taking into account exhaust gas recirculation. The drawback of [7] is the insufficiently fast operation of the piston system with the adjustable engine compression ratio.
The scheme with early intake valve closing has been investigated in [8]. This scheme is based on the intake valve closing at a time when the effective compression ratio reaches a predetermined value. In this case, the limitation of a part of the intake stroke for the mixture admission from the intake manifold is required. Then the valve is closed and the part of the cylinder, which remains at the bottom dead center of the intake stroke, becomes isolated. After the intake valve closing, the charge that appeared in the cylinder first expands to below atmospheric pressure due to the piston movement down during the last part of the intake stroke. Then the charge is compressed again on the compression stroke. The obvious drawback of the proposed scheme is that such early closing of the intake valve reduces admission.
The idea to use the prolonged working fluid expansion cycle instead of the Otto constant-volume cycle, which is the basis for the functioning of the majority of modern spark engines has been implemented in [9]. In this cycle, the expansion ratio was greater than the compression ratio. But such converted engine will have a too complex kinematic diagram of the power mechanism.
The method of implementation of the prolonged expansion cycle due to late intake valve closing to reduce the maximum exhaust gas temperatures has been proposed in [10]. But the converted engine had lower power performance than the base engine.
In [11], the performance of the Opel X17DTL diesel engine converted to run on liquefied propane-butane has been investigated. The reduction of the compression ratio of the converted engine was due to the installation of two additional gaskets between the cylinder head and block. It should be noted that this method of reducing the compression ratio leads to a decrease in the indicated efficiency of the gas engine compared to the base diesel engine.
Thus, the unresolved problem in the conversion of diesel engines to gas is the increased fuel consumption and low indicated efficiency. Therefore, an effective method to ensure a high indicated efficiency of the gas-converted diesel engine by reducing the compression ratio of the base engine due to later intake valve closing has been proposed.
3. The aim and objectives of the study
The aim of the work is to increase the fuel economy of gas-converted diesel engines by reducing the compression ratio of the base diesel engine due to later intake valve closing. To achieve this aim, the following objectives are set: - to perform a theoretical research of changes in the indicated specific fuel consumption for gas-converted diesel
94 ) 201
engines, where the reduction of the compression ratio of the base diesel engine is achieved due to later intake valve closing;
- to carry out an experimental research of changes in the indicated specific fuel consumption for the gas-converted diesel engine, where the reduction of the compression ratio is achieved by the traditional method and the proposed method of later intake valve closing.
4. Calculation of the indicated specific fuel consumption of the converted engine of a new design
The use of prolonged expansion cycles due to late intake valve closing in engines allows:
- reducing the requirements to the anti-knock quality of the gas-air mixture when reducing the actual compression ratio in the engine cylinder;
- implementing the concept of prolonged expansion when reducing the actual compression ratio and, accordingly, increasing the cycle efficiency;
- reducing the intake depression due to the reverse release of the fuel mixture, which will reduce pumping losses in all throttling modes during engine operation.
An objective index of the efficiency of any thermal power plant, including diesel engines converted to gas fuels, is the
indicated efficiency. High values of the latter directly affect the fuel economy and performance of engines. The method to increase the indicated efficiency by reducing the compression ratio of the base diesel engine due to later intake valve closing is shown in Fig. 1. The scheme in Fig. 1, a demonstrates the shift of the intake valve closing point on the indicator diagram. Point a' is the moment of closing of the intake valve of the base diesel engine, point a" is the moment of closing of the intake valve of the converted engine. To shift the angle of closing of the intake valve of the Opel X17DTL diesel engine (Fig. 1, b), cranking of the camshaft drive gear of the valvetrain was carried out (Fig. 1, c).
The calculation of the indicated efficiency n is based on the comparison of the amount of supplied and removed heat. In the simplified calculations, when determining the indicated efficiency of the cycle, only the heat that is supplied and removed in isochoric and isobaric processes is taken into account, and compression and expansion processes are taken without heat exchange, i. e., as adiabatic. In this calculation, when determining the thermal efficiency of the cycle, heat exchange in the compression and expansion strokes was also taken into account. Similar considerations were also used when determining the average theoretical pressure with different methods of organization of engine cycles.
V cm
b c
Fig. 1. Implementation of the method of reducing the compression ratio of the base diesel engine due to later intake valve closing: a — shift of the intake valve closing point a — a" on the indicator diagram; b — Opel Astra diesel engine converted to
the propane-butane mixture; c — camshaft drive gear
a
In the calculations, the authors have made the following assumptions:
- the working fluid is a real gas, the heat capacity of which is constant, i. e. it does not depend on temperature T and pressure P;
- the chemical composition of the working fluid varies, its mass is fixed and depends on the cylinder capacity;
- the duty cycle is closed and reverse, that is, the intake and exhaust periods are not taken into account;
- there is no intake of fresh mixture and release of combustion products;
- there are no energy losses in the cycle, no pumping losses;
- the influence of fuel is taken into account through the lower heating value and the molecular weight of fuel;
- the combustion process is carried out at the end of compression at a constant volume and continues at a constant pressure;
- the cycle is completed at the end of expansion by the process of heat transfer from the working fluid to an external cold source at a constant volume, and then turning the working fluid into the initial state - at a constant pressure;
- the amount of heat transferred is proportional to the mass and combustion heat of a real fuel mixture.
The energy efficiency of engines is estimated by the indicated efficiency n, which is defined as the ratio of heat to perform the useful work Au to the supplied heat Qs obtained as a result of combustion of the fuel-air mixture
n = A = Qs - Q = ! - Q,
(1)
where Qr is the amount of removed heat, kJ.
The amounts of heat supplied to engine cylinders during combustion of the fuel-air mixture and removed (through the exhaust system, cooling system, etc.) heat are determined by temperatures in working processes
The polytropic expansion index is assumed to be «2=1.265, the polytropic compression index - «1=1.364 [12].
Volumes Va", temperatures Ta" and pressures Pa" of the working fluid in the relation (4) at the end of late intake valve closing a" for the converted engine are calculated as follows:
V- =
V -V
2
R
(1 - cos $) + —(! - cos2^)
T. = T
_Vl .v .,
\ a J
\«1 -1
, Pa = Pa
.v., a.
Qs = m£r,rAT,- tc )+m c .(T - T,),
(2)
where R is the crank radius; L is the crank length; 9 is the set angle of late intake valve closing.
Temperatures, pressures and volumes of the working fluid in the relation (4) at the end of the processes of intake a, compression c and expansion b for the converted engine are determined by the formulas:
P T V P
T = a b V = h = V ■£
a Pb ' a p-1 c '
TP
T = P = —
b p"2-1 ' b pnl '
Tc = Tapn -1, Pc = Pap \ Vc = = Va",
p pv
where £ is the geometric compression ratio of the engine, P = Va/Vc; Vh is the converted engine capacity; p is the ratio of the previous expansion of the working fluid, p = |iTz/ (XTC); 5 is the ratio of the next expansion of the working fluid, 8 = pp; £v is the actual compression ratio of the engine, Pv = Va-./ Vc.
Temperatures, pressures and volumes in the relation (4) in the combustion process for the converted engine are calculated as follows:
T = TV=TV=T p T
z Vz. Vc z'z' n '
Q = mcCv c.(Tb -Ta) + mcC(Ta -T,.),
(3)
where Cvw, Cvc are isochoric heat capacities, respectively, of the working mixture and combustion products; CPw w,, CPc are isobaric heat capacities, respectively, of the working mixture and combustion products; mw, mc are the masses of the working mixture and combustion products; Tz', Tc, Tz, Tb, Ta, Ta" are the temperatures, respectively, of the working fluid at the beginning of the previous expansion (Fig. 1), at the end of compression, at the end of the previous expansion, at the end of the next expansion, at the beginning of the geometric compression process, at the beginning of the actual compression process.
Then, from (1)-(3), the thermodynamic efficiency coefficient n\i will be
n = 1 -
mC, (Tb - T) + mC (T - T..)
v.c. V b a ' p.cA a a '
= 1 -
mCv.w.(Tz- - Tc ) + mCp.JTz - T,) Cy.c. ((Tb - Tg ) + «2(Ta - T,.)) CcwiiT- T )+«T - T,))'
(4)
where « is the average polytropic compression index, n2 is the average polytropic expansion index.
P T
P, = V, = Vc p, T
where X is the ratio of pressure increase in the working fluid, | is the coefficient of molecular change of the working mixture.
By determining the indicated efficiency n, the theoretical indicated specific fuel consumption gi, g/kWh is calculated by the formula:
g = 3600/(Hu -n),
where Hu is the lower heating value, MJ/kg.
5. Methods and materials of the experimental research of fuel-economy characteristics of the diesel and converted engine
The aim of the experimental research is to determine the indicated specific fuel consumption of the gas-converted diesel engine. To achieve this aim, the Opel X17DTL diesel engine was converted to run on liquefied propane-butane in the Ivano-Frankivsk National Technical University of Oil and Gas (Ukraine).
For the conversion of the Opel diesel engine to gas fuel, gas cylinder equipment of Italian production was installed. In addition, the compression ratio of the converted engine was reduced, the electronic DIS ignition system of own development was installed and the operation of the converted engine control system was optimized (Fig. 2).
The effective wheel power Ne, kW was determined by the formula
Ne = %
. Me ■ nx 3 104. i
where Me is the effective wheel torque, Nm; nx is the engine speed, min-1; i is the transmission ratio.
The effective wheel torque was determined using the GOSNITI KI-8964 traction-brake stand [11].
The engine mechanical efficiency nen and transmission ntr were determined through the mechanical loss power of the engine Nen and the mechanical loss power of the transmission Ntr.
nen =
N„
Ne + Ne
^ nir = 1 +
Ne + Nen
Nr
Fig. 2. Opel Astra passenger car with the gas-converted diesel engine: a — X17DTL engine; b — the engine with the dismantled head and installed ignition system; c — the head
of the converted engine with plugs and diesel nozzles
The first engine conversion and reduction of compression ratio were made by installing additional gaskets under the cylinder head. The experimental research of fuel-economy and environmental characteristics of the converted engine was carried out [11]. The feature of the second conversion is the reduction of the actual compression ratio due to late intake valve closing. Table 1 shows a brief technical description of the Opel Astra X17DTL converted diesel engine.
The indicated specific fuel consumption g, g/kWh was experimentally determined through the engine mechanical efficiency nen and transmission ntr, hourly fuel consumption G and effective wheel power Ne as follows:
g = G/(Ne ■nen ' ntr ).
The mechanical loss powers were determined when measuring the power consumption of the electric motor, which was interlocked by the retainer with the drive wheel. The mechanical loss power of the transmission Ntr was measured in the neutral position of the gearbox. The mechanical loss power the engine Nen and the mechanical loss power of the transmission Ntr were measured in the highest gear.
The studied materials and equipment used in the experiments are given in detail in [11].
Table 1
Brief technical description of the Opel X17DTL converted diesel engine
No. Parameter Value
Reduction of the actual compression ratio by installing additional gaskets under the head [11] Reduction of the actual compression ratio by late intake valve closing
1 Base engine Diesel, with Bosch EDC 15M injection system
2 Converted engine Gas (propane-butane), with the IFN-TUNG control system
3 Engine capacity, cm3 1,669
4 Maximum power, kW(hp)/rpm, min-1 50(68)/4,400
5 Maximum torque, Nm/rpm, min-1 130/2,000
6 Base diesel engine compression ratio 22.0
7 Base diesel engine intake valve closing angle 30°
8 Converted engine intake valve closing angle 30° 52.5°
9 Converted diesel engine compression ratio 13.1 13.2
10 Thickness of gaskets under the cylinder head 4.2 mm 1.4 mm (nominal)
c
6. Results of the research of the indicated specific fuel consumption of diesel engines converted to gas fuel
To obtain the initial data for comparison in order to evaluate the efficiency of the gas-converted diesel engine, the experimental research of three options of the converted engine was carried out. The first option is the gas-converted diesel engine with the base compression ratio of 22.0. The second option is the gas-converted diesel engine, where the reduction of the compression ratio of the converted engine to 13.1 was due to the installation of two additional gaskets between the cylinder head and block. The third option is the gas-converted diesel engine, where the reduction of the compression ratio of the converted engine to 13.2 was due to the 22.5o later intake valve closing.
Fig. 3 shows the dependencies of the indicated specific fuel consumption and torque of the Opel X17DTL gas-converted diesel engine with the base compression ratio of 22.0 on changes in the crankshaft speed.
g/kW h
240
220
200
180
....... ...
\
— ^ y •........ .
N ------ -
90
Me, Nm
70 60
800
1600
2400
3200 4000 n, min -1
indicated specific fuel consumption was 203 g/kWh. Herewith, the theoretical indicated specific fuel consumption at the nominal mode was 247 g/kWh, while the minimum theoretical value of the indicated specific fuel consumption was 211 g/kWh. Thus, with a decrease in the engine compression ratio due to the installation of additional gaskets, the experimental value of the indicated specific fuel consumption, depending on changes in the crankshaft speed, increased in comparison with the first conversion option in the range from 8.1 to 9.7 %. The theoretical value of the indicated specific fuel consumption increased from 3.9 to 7.9 %.
g/kWh 250
240
220
200
180
130
Me, Nm
110
100
60
50
800
1600
2400
3200
40
4000n, min-1
Fig. 3. Dependencies of the indicated specific fuel consumption and torque of the Opel X17DTL gas-converted diesel engine with the base compression ratio of 22.0 on
changes in the crankshaft speed:--indicated specific
fuel consumption, theoretical dependence;
----— indicated specific fuel consumption, experimental
dependence;.........— engine torque, experimental dependence
According to the results of the experiments, the value of the indicated specific fuel consumption at the nominal mode was 220 g/kWh, the minimum value of the indicated specific fuel consumption was 185 g/kWh. Herewith, the value of the theoretical indicated specific fuel consumption at the nominal mode was 229 g/kWh, while the minimum theoretical value of the indicated specific consumption was 191 g/kWh. It should be noted that in order to avoid knocking of the engine with the basic compression ratio of 22.0 when running on gas, the engine torque did not exceed 70 % or 91 Nm. Obviously, achieving the first option of acceptable power characteristics is not possible in principle.
Fig. 4 shows the dependencies of the indicated specific fuel consumption and torque on changes in the speed of the X17DTL gas-converted diesel engine with the compression ratio of 13.1, where the reduction of the compression ratio of the converted engine was due to the installation of additional gaskets between the cylinder head and block.
As a result of the tests, the characteristics of the gas-converted diesel engine to obtain the original reference base were obtained. According to the results of the experiments, the value of the indicated specific fuel consumption at the nominal mode was 238 g/kWh, the minimum value of the
Fig. 4. Dependencies of the indicated specific fuel consumption and torque of the Opel X17DTL gas-converted diesel engine with the compression ratio of 13.1 due to the installation of additional gaskets on changes in
the crankshaft speed:--indicated specific fuel
consumption, theoretical dependence;----— indicated
specific fuel consumption, experimental dependence; .........— engine torque, experimental dependence
Fig. 5 shows the dependencies of the indicated specific fuel consumption and torque on changes in the speed of the Opel X17DTL gas-converted diesel engine with the compression ratio of 13.2, where the reduction of the compression ratio of the converted engine was due to late intake valve closing.
g/kW h 220
200
180
160 800
_________
/
N \
\
— / / / > /
---N
1600
2400
3200
130
Me, Nm
110 100 60
50 40
4000 n, min -1
Fig. 5. Dependencies of the indicated specific consumption and torque of the Opel X17DTL gas-converted diesel engine with the compression ratio of 13.2 due to late intake valve closing on changes in the crankshaft speed:
--indicated specific consumption, theoretical
dependence;----— indicated specific consumption,
experimental dependence;.........— engine torque,
experimental dependence
According to the results of the experiments, the value of the indicated specific fuel consumption at the nominal mode was 204 g/kWh, the minimum value of the indicated specific consumption was 160 g/kWh. Herewith, the theoretical indicated specific consumption at the nominal mode was 206 g/kWh, while the minimum theoretical value of the indicated specific consumption was 174 g/kWh. Therefore, with a decrease in the engine compression ratio due to late intake valve closing, the experimental value of the indicated specific consumption decreased in comparison with the traditional conversion option in the range of 7.3-8.1 %. The theoretical value of the indicated specific fuel consumption decreased from 8.8 to 10.4 %.
7. Discussion of the results of the research of the indicated specific fuel consumption of diesel engines converted to gas fuel
The research is a continuation of the study of the methods of conversion of diesel engines to gas fuels [11].
The theoretical calculations and experimental research are useful as they demonstrate the effectiveness of the method of optimizing the engine working process due to later intake valve closing. The theoretical and experimental research shows the possibility of a significant reduction, on average, by 11-14 %, of the indicated specific fuel consumption depending on the engine torque. This, in turn, allows expecting a proportional reduction in the operating fuel consumption for gas-converted diesel engines by the proposed method of reducing the compression ratio by later intake valve closing. It should be noted separately that the specified figures of reduction of fuel consumption of engines with the mechanism of later intake valve closing, as the experiments have shown, will be obtained when providing power at the level of the base engine.
As a significant advantage of the research, it can be noted that the theoretical calculations performed are well correlated with the experimental results. Therefore, the research can be recommended and used by specialists of automobile plants and motor-construction enterprises in the design of new and conversion of existing engines to alternative gas fuels. The results obtained allow optimizing the designs of power systems
and valvetrains of gas-converted diesel engines. At the same time, restrictions on the use of the proposed method in terms of the results of fuel economy are obvious. Excessively late intake valve closing may result in reverse fuel emissions into the intake manifold and reduction of the effective compression ratio of the engine. This, in turn, will lead to a deterioration of the combustion process and an increase in fuel consumption.
Therefore, further research into the methods of conversion of diesel engines to gas fuel should be focused on the optimization of valve timing, which is associated with significant experimental difficulties.
8. Conclusions
1. As a result of the theoretical research, the calculations of the theoretical indicated specific fuel consumption for different methods of conversion of diesel engines into gas engines were made. In this case, the calculations showed that for the proposed method of conversion of diesel engines to gas with the 22.5o later intake valve closing, the indicated specific fuel consumption is reduced in the range of 8-10 %.
2. As a result of in vitro experimental research with different methods of conversion of the Opel X17DTL diesel engine, the dependencies of the indicated specific fuel consumption on the crankshaft speed were determined. It was found that using the traditional method of engine derating due to the installation of additional gaskets between the cylinder head and block, the value of the indicated specific fuel consumption increased, on average, by 8-9 %. But with a decrease in the compression ratio of the engine by the proposed method of late intake valve closing, the experimental value of the indicated specific consumption not only did not increase, but decreased, on average, by 7-8 %. The main reason for the reduction of the indicated fuel consumption is that in the engine converted by this method, the expansion ratio will be greater than the compression ratio. And this, accordingly, will increase the indicated efficiency of the engine. In addition, with a decrease in the compression ratio of the engine by this method, the intake depression is decreased, which leads to a reduction in pumping losses in all modes. And this, accordingly, also reduces fuel consumption of the converted engine.
References
1. Main trends of biofuels Production in Ukraine / Panchuk M., Kryshtopa S., Shlapak L., Kryshtopa L., Panchuk A., Yarovyi V., Sladkowski A. // Transport Problems. 2017. Vol. 12, Issue 4. P. 15-26.
2. Experimental Research on Diesel Engine Working on a Mixture of Diesel Fuel and Fusel Oils / Kryshtopa S., Kryshtopa L., Melnyk V., Dolishnii B., Prunko I., Demianchuk Y. // Transport Problems. 2017. Vol. 12, Issue 2. P. 53-63.
3. Grebnev A. V. Influence of the setting angle of the fuel injection advance on the combustion characteristics and the content of nitrogen oxides in the diesel cylinder with turbocharging and intercooling of the charge air 4 ChN 11,0/12,5 // Improving the performance of internal combustion engines. Mater. II All-Russian Sc. Pr. Conf. "Science-Technology-Resourcesaving": Rus. Ac. of Transport. Viatka State Agricultural Academy, 2008. Issue 5. P. 194-197.
4. Zakharchuk O. V. Improvement environmental performance wheel tractor using gas fuel // Visnyk Nats. tekhn. un-tu "KhPI": zb. nauk. pr. Temat. Vyp.: Avtomobile- ta traktorobuduvannia. Kharkiv: NTU "KhPI ". 2014. Issue 10 (1053). P. 27-32.
5. Kaleemuddin S., Rao P. Conversion of diesel engine into spark ignition engine to work with CNG and LPG fuels for meeting new emission norms // Thermal Science. 2010. Vol. 14, Issue 4. P. 913-922. doi: https://doi.org/10.2298/tsci1004913k
6. Saleh H. Effect of variation in LPG composition on emissions and performance in a dual fuel diesel engine // Fuel. 2008. Vol. 87, Issue 13-14. P. 3031-3039. doi: https://doi.org/10.1016/jj.fuel.2008.04.007
7. Nadar K., Reddy R. Combustion and emission characteristics of a dual fuel engine operated with mahua oil and liquefied petroleum gas // Thermal Science. 2008. Vol. 12, Issue 1. P. 115-123. doi: https://doi.org/10.2298/tsci0801115n
8. Experimental investigation and combustion analysis of a direct injection dual-fuel diesel-natural gas engine / Carlucci A. P., de Risi A., Laforgia D., Naccarato F. // Energy. 2008. Vol. 33, Issue 2. P. 256-263. doi: https://doi.org/10.1016/j.energy.2007.06.005
9. Cheenkachorn K., Poompipatpong C., Ho C. G. Performance and emissions of a heavy-duty diesel engine fuelled with diesel and LNG (liquid natural gas) // Energy. 2013. Vol. 53. P. 52-57. doi: https://doi.org/10.1016/j.energy.2013.02.027
10. Elnajjar E., Hamdan M. O., Selim M. Y. E. Experimental investigation of dual engine performance using variable LPG composition fuel // Renewable Energy. 2013. Vol. 56. P. 110-116. doi: https://doi.org/10.1016/j.renene.2012.09.048
11. Research into emissions of nitrogen oxides when converting the diesel engines to alternative fuels / Kryshtopa S., Panchuk M., Dolishnii B., Kryshtopa L., Hnyp M., Skalatska O. // Eastern-European Journal of Enterprise Technologies. 2018. Vol. 1, Issue 10 (91). P. 16-22. doi: https://doi.org/10.15587/1729-4061.2018.124045
12. Numerical study on the combustion process of a biogas spark-ignition engine / Carrera J., Riesco J., Martínez S., Sánchez F., Gallegos A. // Thermal Science. 2013. Vol. 17, Issue 1. P. 241-254. doi: https://doi.org/10.2298/tsci111115152c
-□ □-
Розроблено ттегровану систему тдтримки функщонування теплонасосного енергопостачання на основi прогнозування змти температури м^цевог води. Змта витрати пари холодагента, числа обертiв електродвигуна компресора вiдбуваeться при вимi-рювант температури холодагента на виходi iз конденсатора, тиску випаровування, тиску конденсащг та частоти напруги.
Виконано комплексне математичне моделювання теплонасосног системи, що базуеться на ттегроватй системi тдтримки розряду грунту нарiвнi 10-8°С. Визначено витрату холодагента, потужтсть електродвигуна компресора, напругу, частоту напруги, число обертiв електродвигуна компресора, коефщ1ент продуктивностi теплонасосног системи для встановлених рiвнiв функщонування. Встановлено параметри конвективного тепло-обмту в конденсаторi, постшт часу та коефщгенти математич-них моделей динамжи змши температури мкцевог води, витрати пари холодагента, числа обертiв електродвигуна компресора.
Здобуто функщональну ощнку змти температури мкцевог води в дiапазонi 35-55 °С впродовж опалювального сезону, витрати пари холодагента, числа обертiв електродвигуна компресора. Визначення тдсумковог функциональной тформацп надае можливкть прийма-ти наступт випереджуючiргшення: на тдтримку змти тиску випаровування щодо змти витрати пари холодагента для цифрового управлтня; на тдтримку змти тиску випаровування щодо змти витрати пари холодагента та на змту частоти напруги щодо змти числа обертiв електродвигуна компресора для частотного управлтня. Тому, запропоновано прогнозування змти температури мкце-вог води на основi вимiрювання температури холодагента на виходi iз конденсатора. Саме ця ощнка у стввiдношеннi з вимiрюваним тиском випаровування, входить до складу аналхтичних визначень витрати холодагента та числа обертiв електродвигуна компресора. Здобуття таког ощнки та вимiрювання частоти напруги надае можливкть упереджено впливати на узгодження функщонування зовтшнього та внутршнього контурiв теплонасосног системи як при цифровому, так i частотному управлтт
Ключовi слова: теплонасосна система, частотне управлтня,
цифрове управлтня, тиск випаровування, тиск конденсащг -□ □-
UDC 621.31
| DOI: 10.15587/1729-4061.2018.139473 |
DEVELOPMENT OF ENERGY-SAVING TECHNOLOGY FOR
MAINTAINING THE
FUNCTIONING OF HEAT PUMP POWER SUPPLY
E. Chai kovskaya
PhD, Associate Professor, Senior Researcher Department of Theoretical, general and alternative energy Odessa National Polytechnic University Shevchenka ave., 1, Odessa, Ukraine, 65044 E-mail: eechaikovskaya@gmail.com
1. Introduction
Under conditions for saving natural fuel and decreasing harmful emissions into the atmosphere, heat pump power supply with the use of renewable power sources is gaining further development [1-3]. Thus, for example, the heat pump, using fermented must as a low-potential energy source, is recommended in order to maintain the operation of the biogas plant as a part of a cogeneration power system. Due to additional biogas production, the proposed technology provides an opportunity to increase marketability of a biogas plant and decrease the cost value of production of electricity and heat in the range of 20-30 % [3].
The outside air as a low-potential power source is available in heat pump power supply, but the change of air temperature in a wide range makes it difficult to maintain functioning of heat pump systems [1].
A heating system of the soil-water type requires the construction of special soil heat exchangers for heating the brine - 30 % ethylene glycol solution that is fed to the evaporator of a heat pump [2]. A vertical heat exchanger occupies a smaller area than a horizontal one but requires additional capital investments for drilling wells. It is possible to ensure heat extraction within 30-100 W per one meter of the length of a vertical heat exchanger depending on a soil type at the depth of 40-150 m, where the soil temperature
