5.1.7) The value of CGtl is the superposition particle velocity at the first mesh point in
Chapter 5 Computer Modelling of Engines
5.2.3 The chainsaw engine at full throttle
The data for the model to be analyzed by Prog.5.1 is given in Figs. 5.3 and 5.4(a).
In Fig. 6.1 there is a sketch of port areas related to crankshaft angle, as part of a general discussion on the topic, however the actual geometry used to create Fig. 6.1 is that for the chainsaw engine under discussion. As a further aid to physical visualization, Fig. 6.2 shows a computer generated sketch of this same chainsaw engine cylinder. The data could be regarded as being virtually a universal chainsaw engine design, for there are many such engines in production with very similar porting and exhaust systems. The hallmark of these small high-speed engines, used for such tools as chainsaws, brush cutters and snow blowers, is a small, compact silencer with a limited volume so that the bulk of the entire device does not make it unwieldy. The porting layout of many of these engines is remarkably common, with a piston controlled inlet port, usually two transfer ports and a single exhaust port. The main aim of the designer is to attain the requisite power output at as low a sound and vibration level as possible, for most of these tools are hand-held. Fuel consumption is also an important criterion, for although the engines are small in terms of both cylinder capacity and total power output, they are often used profes- sionally in excess of 1000 hours per annum. In other words, the total fuel consumed
can be considerable and a reduction in the specific fuel consumption rate is noticeable. This is particularly true as most of these hand-held tools have quite small capacity fuel tanks, so the operator tends to observe, and mentally record, the time interval between refueling.
There is one curiosity in the data bank in Fig. 5.3 which may raise a query in the minds of some readers. The engine data apparently specifies that the engine is running at 3/4 throttle, rather than at full throttle. The reason for this is that such engines use diaphragm carburetors with a very pronounced venturi giving a restriction ratio of 25% to the carburetor flow diameter, even when the butterfly throttle valve is wide open! Such are the mathematical subterfuges to which the programmer must resort when it is necessary to accurately describe some particular engine geometry.
It will be seen from the data table that, although the port timings are not dissimilar to the QUB 400 engine, the exhaust box silencer is small and restrictive. As further acquaintance with the technical jargon, silencers are often referred to in automotive literature as "mufflers." The volume of this box is merely five times larger than the cylinder capacity, and the outlet pipe area from the box is just 40% of that of the exhaust downpipe connecting the box silencer to the exhaust port. The pipes at either side of the muffler are also very short, each being 90 mm.
The output of the computer calculation using Prog.5.1 is presented in Fig. 5.14 for a crankcase compression ratio of 1.4. It will be seen that the restrictive exhaust box has a profound influence on the cylinder and exhaust pressure diagrams. The build-up of exhaust back-pressure during the scavenge phase opposes the crankcase blowdown with the result that the crankcase pressure is still above the atmospheric pressure by exhaust port closure. This reduces the strength of the suction to half of the crankcase pumping action which translates into a low delivery ratio of 0.487.
Weaker suction pulses provide lesser intake ramming behavior which compounds the problem. As the cylinder pressure remains high during the scavenge period, the expansion of the scavenging volume flow is also limited and the SRv value is only marginally higher than the scavenge ratio level, i.e., SRv is 0.614 for a SR level of 0.537. The resulting scavenging efficiency is correspondingly low at 0.52. Never- theless, the engine produces 3.66 bar BMEP at 7500 rpm and a power output of 4.6 kW, with a commendable BSFC of 444 g/kWh. This would be regarded as a normal performance characteristic for such an engine.
With the restrictive exhaust system having been highlighted as the limiter on engine performance, the logical design option is to examine an engine design with one which is less restrictive. If that is not a viable design option, due to both engine bulk and noise legislation limitations, then an alternative is to attempt to incorporate a stronger air pumping action. Thus a data change is made to give a crankcase with a compression ratio of 1.5, and the calculation is repeated with the result shown in Fig. 5.15. The delivery ratio is considerably higher at 0.562, the scavenging efficiency is increased to 0.564, the power rises to 5.4 kW, and the BMEP to 4.3 bar.
Although the pumping MEP has risen to 0.29 bar from 0.21 bar, the increase of IMEP from 4.18 to 4.90 bar more than compensates for the greater pumping work.
Chapter 5 - Computer Modelling of Engines
2.0 P.
A T II 1.5
1 o
cv
-
. EXHA
- /
.INDER
J S T> ^ ~
Y A M 1 TYPE SCAVENGING
CRANKCASE
NOTE S M A L L EXHAUST BOX INCREASES EXHAUST PRESSURE AND OPPOSES SCAVENGING
/
EP0 TP0 BDC TPC EPC CRANK-ANGLE from EXHAUST PORT OPENING to EXHAUST PORT CLOSING ENGINE SPEED, rpm= 7500 DELIVERY RATI0=0.487
POWER, kW= 4.6 BSFC,kg/kWh=0.444 BMEP, bar=
IMEP, bar= 4.18 PMEP, bar= 0.21 FMEP, bar=
PEAK CYLINDER PRESS, bar= 33.2 and TEMP., K=2337. at deg.
5
0
7
-
^^
CRANI
EPC
DELIVERY RATIO
\ j
:CASE s^
IPO
/ cVlND^R— —
WEAKER PUMPING ^ ~ - ^ GIVES LOWER SUCTION
PRESSURE
TDC IPC 3.66 0.31 VTDC=13.9
SJLET/
0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1 0.0 E F F.
EP0 0
CRANK-ANGLE from EXHAUST PORT CLOSING to EXHAUST PORT OPENING SCAVENGING EFF.=0.522 TRAPPING EFF.=0.562 CHARGING EFF.=0.274 SCAV. RATI0(mass)=0.537 SCAV. RAT10(vol)=0.614 EX. FLOW RATI0=0.532
Fig. 5.14 Predicted performance characteristics of the
chainsaw engine with a CCR of 1.4.
The Basic Design of Two-Stroke Engines
The result is a reduction of brake specific fuel consumption from the original design value at 444 g/kWh to 435 g/kWh. The stronger pumping action on both the compression and suction phases of the crankcase operation due to the higher CCR is clearly evident by comparing Fig. 5.14 with Fig. 5.15.
The overall performance characteristics of the chainsaw engine at full throttle over a speed range of 5000 to 9000 rev/min is calculated by altering the data input for engine speed in the opening data block in Fig. 5.3. This has been done at 1000 rpm intervals and the result for both crankcase compression ratio values of 1.4 and 1.5 is presented in Fig. 5.16. The output data for power, BMEP, BSFC, and delivery ratio at each engine speed is presented in graphical form. The actual values plotted at each calculation speed are the average of the performance characteristics determined for the third, fourth and fifth engine cycles. The variation of perform- ance characteristics with each succeeding cycle is small for an engine of the industrial type with an untuned exhaust system; in Sect. 5.4 a quite different form of cyclic behavior will be demonstrated. In the reality of a firing engine, as in a computer calculation, there are cyclic variations of all the performance character- istics but this is particularly true for the combustion diagram.
In Fig. 5.16, as found at the particular speed of 7500 rpm plotted in Figs. 5.14 and 5.15, the higher crankcase compression ratio of 1.5 performs better across the entire engine speed range. The peak power output at 8000 rpm has been raised by 18% with a marginal improvement of 4% in specific fuel consumption. It reinforces the opinion that the crankcase pumping action is one of the most significant design para- meters for a high-speed industrial engine of this type(5.1).
The almost linear fall of DR with the lower crankcase compression ratio of 1.4 results in a 20% drop in power at 9000 rpm. At 5000 rpm there is very little difference in performance between the two variations of the engine, emphasizing the point that it is for the design of a high-speed engine that attention must be paid to the optimization of the crankcase pumping action. Equally, it could be that the designer may have a particular reason to have the power level fall off sharply at high speed, as there may be some legal safety requirement to be met regarding the upper speed which can be attained by an operator using the device. In this case also, the designer can tailor the CCR level to meet such a need, for the sharp fall in power from 8000 to 9000 rpm in the lower CCR example is evidence that this is technically possible.
5.2.4 Concluding discussion on Prog.5.1, "ENGINE MODEL NO.l"
It is clear that this model allows the designer to produce predicted engine
performance characteristics for a parameter change which could be as major as crankcase compression ratio, or as minor as a radius variation in the top corners of a port. By such means, the designer can tailor the performance characteristics of the engine to meet the market needs of the device and thus eliminate many of the ex- perimental "cut and try" techniques so common in industry. This does not mean that the unsteady gas-dynamic program is so accurate as to replace experimentation, such is definitely not the case, but it does permit a more rapid understanding of the
Chapter 5 - Computer Modelling of Engines
2.0 P.
A T M
1.5
1.0
TP0 BDC TPC EPC CRANK-ANGLE from EXHAUST PORT OPENING to EXHAUST PORT CLOSING ENGINE SPEED, rpm= 7500 DELIVERY RATI0=0.562
POWER, kW= 5.4 IMEP, bar= 4.90 PEAK CYLINDER PRESS
BSFC,kg/kWh=:0.435 PMEP, bar= 0.29 barr 36.6 and TEMP.
BMEP, bar= 4.30 FtlEP, bar= 0.31 K=2466. at deg.ATDC=12.3
0.8 1.5
P.
A T M
1.0
0.7 IPO 'TDC IPC EP0
CRANK-ANGLE from EXHAUST PORT CLOSING to EXHAUST PORT OPENING SCAVENGING EFF =0.564 TRAPPING EFF.=0.550 CHARGING EFF.=0.309 SCAV. RATI0(mass)=0.602 SCAV. RATIO(vol)=0.669 EX. FLOW RATI0=0.600
Fig. 5.15 Predicted performance characteristics of the
chainsaw engine with a CCR of 1.5.
potential behavior of any engine undergoing design or development. It is from such understanding that design and development problems are solved. In the design case, the problem can often be eliminated before the first drawings are made, casting dies purchased, or a prototype machined and constructed. In the experimental situation, the simulation often provides the answer to a baffling problem; in the chainsaw example just discussed, the opening gambit of the experimenter might be to test a series of engines with longer inlet timings in a fruitless attempt to raise the delivery ratio. In either situation, the potential for savings of time, effort and money is considerable.
One word of caution is required regarding scavenging behavior. Within the calculation packages, the designer can allow the "paper" engine to be scavenged as well as a uniflow engine or as badly as a classic deflector piston engine. The only means of ensuring that the actual engine has a scavenging system as good as can be attained, within the design limitations of the engine in question, is to conduct tests using the theoretical or experimental techniques discussed in Chapter 3.
A general conclusion can be made regarding the design of engines with untuned exhaust systems, that in the absence of any effective engine tuning by pressure waves, apart from the ramming effect in the intake system, the BMEP or torque characteristic with speed is effectively controlled by the shape of the delivery ratio curve and the quality of the scavenging process.
5.3 Using Prog.5.2, "ENGINE MODEL No.2"
From Fig. 5.2(b), it can be seen that this computer model contains a tuned exhaust pipe of the type commonly seen on high-performance racing motorcycle engines;
it is also fitted to high-performance road-going motorcycles with silencers of varying degrees of effectiveness attached to the ends of the tuned pipes. Further discussion on this particular topic can be found in Chapter 8. As remarked in Sect.
1.3.2, this form of tuned exhaust pipe was successful in the late 1950's on the MZ racing motorcycles from Zschopau in East Germany. The author does not know who invented this type of exhaust system, colloquially referred to as an "expansion chamber," for racing two-stroke engines, but it is known that Walter Kaaden, the Chief Engineer of MZ, played a major role in its early development and testing( 1.1).
At the same time, the author has clear memories of seeing a 125 cc racing Montesa from Spain which had a very similar type of exhaust system, and this was several years before the advent of the MZ machines. The application of the expansion chamber on a QUB 500 cc single-cylinder racing machine(5.4)(5.5) is shown in Plate 5.1.
The result of the use of this type of tuned exhaust system, allied to a disc valve inlet system with its asymmetrical inlet port timing, was that the two-stroke engine employed for motorcycle racing produced more specific power output than the four- stroke engine. This situation persists to this day in all forms of motorcycle racing and competition.
At QUB there has been, and still exists, considerable research in the understand-
ing, design and development of tuned exhaust systems for all kinds of two-stroke
engines(1.13)(2.6)(2.15)(3.7)(4.9)(4.10)(5.2-5.5). Therefore, the computer model
Plate 5.1 The QUB 500 single-cylinder 68 bhp engine with the expansion chamber exhaust slung underneath the motorcycle (photo by Rowland White).
represented by Prog.5.2 is the result of many years of theoretical research allied to experimentation with high specific output engines. Although Prog.5.2 does not contain induction system variations to handle disc valve or reed valve induction systems, the program permits the designer to model all of the engine porting and exhaust geometry of a high-performance engine. There will be discussion in a later chapter on the design of racing engines and the interpretation of calculated data from a power unit with piston controlled induction for the design of reed valves and disc valves. It has been amply demonstrated by Blair et al(5.9) and Fleck et al(4.10)( 1.13) that engine modelling can incorporate reed valve induction, and predict both the reed valve motion and the delivery ratio. The incorporation of disc valve induction into a computer program like Prog.5.2 is quite straightforward, as can be observed from the relevant design discussions in Chapter 6.
It is rather interesting to consider that in Sect. 5.2.3, dealing with the chainsaw type of engine, adverse comment is passed on the debilitating effect on engine performance posed by a restrictive exhaust system. Here, if one interprets the data in a simple fashion for the expansion chamber proposed in Fig. 5.4(b) for the 125 Grand Prix engine, the exhaust pipe at the exhaust port has a diameter of 38 mm and an outlet diameter to the atmosphere of just 20 mm. There are those who would logically consider that to be "restrictive"! The following discussion will reveal that this area restriction at outlet is vital to the attainment of a high specific power output.
Before commencing to discuss the behavior of expansion chamber exhaust systems, it is important to describe the format of the computer screen output from
Chapter 5 - Computer Modelling of Engines Prog.5.2. The program will produce line-print output of all input data and calcula- tion output, including all of the screen graphs. The latter is the principal output of the analysis and is the basis of Figs. 5.17-5.20 regarding the Husqvarna MK4 engine.
Taking Fig. 5.17 as an example of the screen output from Prog.5.2 while it is running, the x-axis is the crankshaft angle commencing at exhaust port opening, EO, for one complete cycle; exhaust port closure is marked as EC. The other timing events are for the transfer port, i.e., opening and closing is TO and TC, respectively, and for the inlet port, i.e., opening and closing is 10 and IC, respectively. The markers for top and bottom dead center are self-explanatory as TDC and BDC. The diagram on the computer screen is refreshed at the conclusion of each cycle. There are two sets of pressure ratio-time histories being written simultaneously. The top one with its scale at the left is for cylinder pressure, exhaust pipe pressure (that value next to the exhaust port), and crankcase pressure. To prevent a programming problem, the maximum cylinder pressure plotted is at 2.0 atm. The bottom set, with its scale of pressure ratio at the right, is for the crankcase pressure and for the inlet pipe pressure in the pipe next to the intake port. Unlike Prog.5.1, the cylinder pressure diagram during combustion does not appear, because the major interest is in the tuning effect of the exhaust system and its control over engine power and torque. This latter information is presented at the base of the diagram at bdc on each cycle, as seen in Fig. 5.17 for the fifth engine cycle in that example. There, the information includes the peak cycle maximum pressure, temperature and crank angle location and includes that for SE, TE, and CE on each cycle as for Prog.5.1.
5.3.1 Analysis of data for a Husqvarna motorcycle engine using Prog.5.2