JPH0528352Y2 - - Google Patents

Description

[Detailed description of the invention] [Technical field of the invention] The present invention relates to an exhaust brake device that increases engine braking force by blocking an exhaust passage, and particularly relates to an exhaust brake device that increases the braking force. Regarding.

[Prior Art] As is well known, in vehicles such as freight cars, so-called exhaust brake devices are widely used to increase the engine braking force by blocking the exhaust passage of the engine.

FIG. 13 is a schematic diagram of an engine equipped with a conventional exhaust brake device 1, and FIG. 14 is a schematic diagram showing a state in which the exhaust brake device 1 is in operation. As illustrated, the exhaust brake device 1 includes:
Exhaust port #1 of each cylinder #1, #3, #5...
An exhaust brake valve 4 for closing the exhaust passage 3 is provided in the straight pipe portion 2a of the exhaust manifold 2 which is connected to the exhaust manifolds 2a, #3a, and #5a and collects them. The operating principle of the exhaust brake will be explained with reference to FIG. 14. The #3 cylinder is in the exhaust stroke and is pushing out exhaust gas. Normally, the exhaust gas should be released into the atmosphere through the exhaust manifold 2, but since the exhaust manifold 2 is closed by the exhaust brake valve 4, the exhaust gas is filled into the exhaust manifold 2. This is each cylinder #1,
By repeating steps #3, #5, etc., the pressure inside the exhaust passage 3 increases and eventually stabilizes at a certain constant pressure. That is, the pressure inside the exhaust passage 3 is the same as that of the exhaust valves 5 #1 and 5 of each cylinder #1, #3, and #5.
It acts on 5#3 and 5#5 and tries to push them down. However, when the cylinder #5 is in the compression stroke (the same applies to the expansion stroke), the pressure inside the cylinder is higher than the pressure inside the exhaust passage 3, and the exhaust valve 5 #5 is never opened. On the other hand, when the #1 cylinder is in the intake stroke, the closing force of the exhaust valve 5 #1 is only the set force of the valve spring 6, and if the pushing down force due to the pressure in the exhaust pipe exceeds this set force, the exhaust valve closes. 5#1 will be released. Then, the exhaust gas in the exhaust passage 3 flows into the cylinder #1, which causes the exhaust gas in the exhaust passage 3 to flow into the cylinder #1.
When the internal pressure decreases and its downward force becomes lower than the set force of the valve spring 6, the exhaust valve 5#1 closes. This series of operations is repeated for each cylinder #1, #3, #5, . . . , and the pressure in the exhaust passage 3 becomes approximately constant.

On the other hand, the engine tries to overcome the constant pressure in the exhaust passage 3 and exhaust the air, and this energy becomes a braking force, increasing the engine braking force.

In other words, the increased amount becomes the exhaust braking force. Therefore, if the engine is the same, the exhaust braking force will be determined by the pressure in the exhaust passage 3.

[Problems to be solved by the invention] By the way, the present applicant has proposed that the exhaust braking force is not simply determined by the pressure in the exhaust passage 3, but also by the influence of exhaust pulsation occurring in the exhaust passage 3. Through various experiments, we learned that this is the case. In other words, if the peak of the exhaust pulsation (maximum peak pressure) acts on the exhaust valve 5#1 during the intake stroke, the exhaust valve 5#1 will reach the desired valve opening setting pressure when the average value of the internal pressure of the exhaust passage 3 It was learned that the exhaust brake force would be released before reaching the desired exhaust brake force, making it impossible to obtain the desired exhaust brake force.

As a result, the applicant has previously
In the publication, we proposed an ``exhaust brake device for automobiles'' that can secure exhaust braking force by preventing interference of exhaust pulsations between cylinders.

However, as the trend in recent years has been to increase the output of engines using turbochargers and the like, the exhaust braking force has begun to be insufficient to compensate for the increase in output. In other words, since the exhaust braking force depends on the engine displacement itself, if the engine output is increased by installing a turbo supercharger etc., the output will inevitably increase without increasing the exhaust braking force. The braking force for this part will depend on the foot brake. For this reason, a problem arises in that the foot brake wears out quickly, and it has become desirable to further strengthen the exhaust braking force in order to improve its durability.

[Purpose of the invention] The present invention was devised in consideration of the above-mentioned circumstances, and its purpose is to provide an exhaust brake device that can increase exhaust braking force by utilizing exhaust pulsation.

[Summary of the invention] In order to achieve the above object, the present invention collects each exhaust port of a multi-cylinder internal combustion engine in a divided manifold for each cylinder where exhaust interference is difficult to occur, and connects each divided manifold to a collecting pipe through a straight pipe part. In an exhaust brake system that has an exhaust brake valve that opens and closes the pipes downstream of the collecting pipe, the length of the straight pipe section of the split manifold is such that when the exhaust brake valve is closed during high-speed engine rotation, A trough with low gas pressure due to exhaust pulsation that pulsates in the exhaust passage between the exhaust brake valve and the exhaust valve of each cylinder is formed to a length that acts on the exhaust valve of the cylinder during the intake stroke. It is characterized by

According to this configuration, when the exhaust brake is activated when the engine is rotating at high speed, the valley of the exhaust pulsation (low gas pressure) acts on the exhaust valve of the cylinder in the intake stroke, so that the exhaust valve opens. As a result, the average pressure in the exhaust passage increases and the exhaust braking force increases.

[Embodiment] Hereinafter, a preferred embodiment of the present invention will be described in detail using a V-type 10-cylinder engine as an example with reference to the accompanying drawings.

In FIG. 1, reference numeral 10 is a V-type 10-cylinder multi-cylinder engine, and five cylinders #2, #4, #6, #8, and #10 are arranged in the upper left bank 11 shown in the figure and in the lower right bank 12. #1, #3, #5, #7,
Five cylinders #9 are arranged respectively. The firing order of each cylinder #1, #2, ... #9, #10 is #1 → #8 → #7 as shown in Figure 2.
→#6→#5→#4→#3→#10→#9→#2
The cylinders of left and right banks 11 and 12 are ignited alternately. Then, in the right bank 12, the cylinders are in the order of #1 → #7 → #5 → #3 → #9, and in the left bank 11, the cylinders are #2 → #8 → #6 →
The cylinders are ignited in the order of #4→#10, and the ignition order of each cylinder is symmetrical between the right bank 12 and the left bank 11. Therefore, as shown in Figure 1, the exhaust system is divided into right bank 12 and left bank 11.
are provided separately and independently, and the right bank 12
side exhaust system 13a and left bank 11 side exhaust system 13
The respective exhaust systems 13a, 13b are formed symmetrically with respect to
An exhaust brake device is provided at 13b.

Hereinafter, the exhaust brake device 14 of the exhaust system 13a on the right bank 12 side will be explained. As shown in the figure, the exhaust brake device 14 includes cylinders #1,
Exhaust pipe 16 forming an exhaust passage 15 for collecting and flowing exhaust gases from #3, #5, #7, and #9.
The exhaust pipe 16 is provided with each cylinder #1,
It is composed of an exhaust brake valve 21 that closes the exhaust passage 15 after the exhaust gases from #3, #5, #7, and #9 are collected.

The exhaust manifold 17 of the exhaust pipe 16 includes a plurality of (two in this embodiment) divided manifolds 17a,
17b, and branch pipes 18#1, 18#3, 1 of each divided manifold 17a, 17b
8#5, 18#7, 18#9 are connected to exhaust ports #1a, #3a, #5a and #7a, #9a of cylinders #1, #3, #5 and #7, #9, respectively. Connected. The straight pipe portions 19a, 19b of each of the divided manifolds 17a, 17b are further collected in a collecting pipe 20, and an exhaust brake valve 21 is provided downstream of the collecting part of the collecting pipe 20.

Each branch pipe 18#1, 18#3, 18#5, 18
#7, 18 and #9 are connected to the exhaust pulsation valley (minimum peak pressure) from each cylinder #1, #3, #5, #7, #9 that is connected to the other cylinders during the intake stroke. The divided manifolds 17 are grouped into those that can easily act on the exhaust valves (not shown).
a, 17b are integrally molded. In other words, cylinders located in the first half of the intake stroke and cylinders located in the second half of the exhaust stroke, which are likely to cause exhaust interference, should not be grouped together as much as possible. Avoid acting on the exhaust valve. In other words, the second
In the figure, cylinder #9 is grouped with cylinder #1, cylinder #1 is placed with cylinder #7, cylinder #7 is placed with cylinder #5, and so on, so that cylinders #9 and #1 are not grouped together as much as possible. Therefore, in this embodiment, there are two cylinder groups: #1, #5, #3 cylinder group (ideally, it is better to separate #5 cylinder and #3 cylinder) and #7, #9 cylinder group. Split manifolds 17a, 17b into groups
is formed. In addition, each divided manifold 17
a, 17b are further collected in a collecting pipe 20, and all cylinders #1, #3, #5, #7,
An exhaust brake valve 21 is provided downstream of the merging portion immediately after the exhaust gases from #9 are merged.

The graphs shown in FIGS. 3 to 6 show the exhaust passages 15 of the exhaust manifold 17 formed separately as described above.
This is an estimate of the pressure fluctuations (pressure waves) within.
In estimating this pressure wave, the following five items are assumed as the conditions.

The period during which the pressure inside the exhaust passage 15 is increased during the exhaust stroke of each cylinder #1, #3, #5, #7, and #9 is 180° due to excessive cranking.

The basic pressure waves generated in each cylinder #1, #3, #5, #7, #9 are generated by the exhaust brake valve 21 (i.e.
reflected at the closed end of the exhaust passage 15).

When the pressure wave passes through the confluence part (collecting pipe 20) of the divided manifolds 17a and 17b, the attenuation rate is
Attenuates by 0.8. That is, between the same divided manifolds 17a (or 17b), each branch pipe 18
#1, 18#3, 18#5 or 18#7,1
It is assumed that no attenuation occurs at the junction of 8 and 9.

The pressure wave is reflected four times.

The exhaust temperature is 400℃ and the propagation speed is the speed of sound.

First, based on these assumed conditions, the pressure waves generated when the exhaust brake valve 21 is closed during the exhaust stroke of each cylinder #1, #3, #5, #7, and #9 are estimated, and the graph in Figure 3 is obtained. become.

That is, the triangle indicated by a in this graph is the fundamental pressure wave directly generated by the exhaust stroke, and b, c, and d are the fundamental pressure waves generated after their reflection, and are synthesized to generate waves for each cylinder #1 and #. Pressure waves A generated in the exhaust strokes 3, #5, #7, and #9 are obtained. Similarly, when the pressure wave A is reflected by the exhaust brake valve 21 and propagated to the other divided manifold (for example, the pressure wave A generated on the divided manifold 17a side
When the pressure wave B propagates to the other divided manifold 17b side), the pressure wave B becomes as shown in the graph of FIG.

Therefore, these pressure waves A and B are synthesized in accordance with the opening/closing timing of the exhaust valves of each cylinder #1, #3, #5, #7, #9, and also taking into consideration the propagation time. Then, the exhaust port #1a of each cylinder,
The pressure waves generated at #3a, #5a, #7a, and #9a can be estimated. The graph in Figure 5 is an estimated graph of the pressure wave C generated at the exhaust port #1a section of the #1 cylinder, and the graph in Figure 6 is an estimated graph of the pressure wave D generated at the exhaust port #7a section of the #7 cylinder. be.

In the graph of Figure 5, the absolute standards are the opening/closing timing of each intake/exhaust valve and the pressure wave A#1 from the #1 cylinder, and the other pressure waves A#5, A
#3, A#7, A#9 are branch pipes 18, respectively.
#1, 18#3, 18#5, 18#7, 18#9
length and each divided manifold 17a, 17b
By determining the lengths of the straight pipe portions 19a and 19b in consideration of the propagation time of each pressure wave, the starting point can be adjusted arbitrarily to some extent. The important point in this graph is that when the intake valve of #1 cylinder is opened (during the intake stroke), the trough (minimum peak pressure) of exhaust pulsation (pressure wave c) at the exhaust port #1a is 22.
At least the peak (maximum peak pressure) 23 part of the exhaust pulsation (pressure wave c) is made to act immediately before the intake valve closes after the latter half of the intake stroke. These things can also be said about other cylinders #3 (in this example, cylinder #5 is an exception), and during the intake stroke of cylinder #3, there is a valley of exhaust pulsation at the exhaust port #3a. Let it work.

Further, the above is exactly the same for each cylinder #7 and #9 on the other divided manifold 17b side shown in the graph of FIG.

That is, the exhaust manifold 17 connects each branch pipe 18
#1, 18#3, 18#5, 18#7, 18#9
the length of the section, and the method of grouping the divided manifolds 17a and 17b to group them;
By appropriately setting the lengths of the straight pipe portions 19a and 19b of the divided manifolds 17a and 17b,
During the intake stroke of each cylinder #1, #3, #5, #7, #9, its exhaust port #1a, #3a, #5a,
The parts #7a and #9a (ie, each exhaust valve) are formed separately so that the valley of exhaust pulsation acts on them.

The graph in Figure 7 shows the pressure waveform (solid line) of the exhaust pulsation of the actual measured value at the exhaust port in a cylinder in one bank of a V-type 8-cylinder engine, and the exhaust pulsation pressure waveform (solid line) obtained by the above estimation method. This is a comparison with the estimated pressure waveform (broken line). As a result, the phases of high-pressure waves and low-pressure waves in the measured waveform and estimated waveform qualitatively match, and estimated waveforms C and D in Figures 5 and 6 are evaluated to be sufficiently valid. can.

Next, the operation of the present invention will be explained.

As shown in the graphs of FIGS. 5 and 6, the exhaust brake device 14 of the present invention allows the exhaust ports #1a, #3a, and #5
The trough 22 of the exhaust pulsation acts on the a, #7a, and #9a portions during the intake stroke of each cylinder, and even if the crest 23 acts, the intake valve closes in the latter half of the intake stroke. It is designed to work immediately before. In particular, in FIG. 6, when the intake valve of the #7 cylinder is opened and when the intake valve of the #9 cylinder is opened, the exhaust pulsation peak (maximum peak pressure) 23 does not act at all. Therefore, the maximum pressure at the valleys 22 and 22 of the exhaust pulsation that acts on the exhaust ports #7a and #9a during the intake stroke is:
until the exhaust valve opening pressure (determined by the valve spring set force) is exceeded.
The exhaust valve is maintained closed, and as a result, the average pressure within the exhaust ports #7a and #9a (exhaust passage 15) is maintained high and the exhaust braking force is increased. In addition, in Fig. 5, 23 peaks of exhaust pulsation (maximum peak pressure) are acting on the exhaust port portion of the cylinder during the intake stroke, but with the exception of the intake stroke of #5 cylinder, #1 cylinder and #3 cylinder In both cylinders, peaks 23 and 23 of the exhaust pulsation act on the exhaust ports #1a and #3a immediately before the intake valve closes in the latter half of the intake stroke. Therefore, even if the exhaust valve opens near the end of the intake stroke, the intake valve is immediately closed, so that exhaust ports #1a and #3a (exhaust passage 15)
The amount of exhaust gas blown back from the engine to the intake port is suppressed as low as possible, and as a result, the pressure within the exhaust passage 15 is maintained high and the exhaust braking force is increased. Furthermore, in this case, high-pressure exhaust gas flows into the cylinder, so the pumping load in the next compression stroke is also increased and the exhaust braking force is increased.

Furthermore, as shown in Fig. 8, the divided manifold 1
7c, 17d, and 17e are provided, and each cylinder #1 of the left and right banks 11, 12 is
#3, #5, #7, #9 and #2, #4, #6,
#8 and #10 may each be divided into three or more groups. When dividing into three groups, ideally, for example, on the right bank 12 side, cylinders #1 and #3, cylinders #5, and cylinders #7 and #9
It is better to divide them into cylinder groups, but as shown in the figure, there are #1 cylinder, #3 and #5 cylinder groups, and #7 and #9 cylinder groups.
Even if it is divided into cylinder groups, there is an effect of increasing the exhaust braking force to some extent.

The graph in FIG. 9 shows the exhaust brake device 14 having the two-split exhaust manifold 17 shown in FIG. 1, the exhaust brake device 14a having the three-split exhaust manifold shown in FIG. An exhaust brake device 1 having an undivided standard exhaust manifold 2 shown in FIG. These are experimental results comparing the actual measured values of each exhaust brake force. As clearly shown in this graph, the engine speed of the 2-split type and 3-split type is lower than that of the standard type.
It can be seen that the exhaust braking force increases as the engine speed increases above 1400 rpm. In other words, the exhaust braking force of the 2-split type and 3-split type is lower than that of the standard type.
It increases as the engine speed increases to 1500rpm, 2000rpm, 2500rpm, and 3000rpm. That is, the lengths of the straight pipe portions 19a and 19b in the two-split type and the three-split type are tuned to match the high rotation range (1400 rpm or more) of the engine.

The graph in FIG. 10 shows the exhaust braking force and These are the results of experimental measurements comparing the pressure inside the exhaust pipe (pressure inside the exhaust passage). From this result, it can be seen that a higher exhaust braking force can be obtained by combining the divided manifolds 17a and 17b into one exhaust manifold 17. Note that 25 shown in the graph indicates the exhaust braking force of the standard type exhaust braking device.

FIG. 11 shows each divided manifold 17a,
A joint pipe is provided between the manifold 17b and its collecting pipe 20 to separate the respective divided manifolds 17a and 17b.
These are the experimental results of measuring the change in exhaust braking force when the lengths of the straight pipe portions 19a and 19b were changed. The joint pipes include those whose length L is L ₁ = 440 mm, L ₂ = 32 mm, L ₃ = 120 mm, and
Measurements were made using four types of L ₄ =0 mm (that is, no joint pipe). It can be seen from the graph of the experimental results that when the length of the straight pipe portions 19a, 19b of the divided manifolds 17a, 17b is increased, the exhaust braking force increases as the engine speed increases. In addition, exhaust brake valve 2
If we measure only the engine braking force without blocking 1, as shown at 26 in the graph,
Straight pipe portion 19 of split manifolds 17a, 17b
There was no effect of the lengths of a and 19b. Furthermore, although it is not clearly shown in the graph of FIG.
It has been found that there is a correlation between the lengths of 9a and 19b and the engine speed range in which the exhaust braking force begins to increase, and that as the lengths become longer, the exhaust braking force tends to increase from a lower rotation range.

FIG. 12 shows each divided manifold 1 in the three-divided exhaust brake device 14a shown in FIG.
These are the experimental results for measuring the exhaust braking force when the shape of the collecting pipe 20 portions 7c, 17d, and 17e are changed in shape. As shown in the schematic diagram in the graph, Fig. 1 is a model in which the divided manifolds 17c, 17d, and 17e are not assembled, but each is provided with an exhaust brake valve 21, and there is no exhaust interference between the divided manifolds 17c, 17d, and 17e. If it doesn't happen. B shows the case where each divided manifold is connected with a pipe of the same inner diameter and is closed, C shows the case where 20 sections of the collecting pipe are made into a chamber to increase the capacity (approximately 1000 cm ³ ), and D shows the case where the capacity of the chamber is made small (about 64cm ³ ). It can be seen from the graph of this experimental result that the performance of B and D is almost the same, the performance of C is slightly lower than that of B and D, and the performance of A is even worse than C, and especially in the high speed range, it is about about the same as that of B and D. 32% decrease. In other words, although there is no difference depending on the shape of the merging portion 20a, it is understood that the smaller the volume of the merging portion 20a, the better.

[Effects of the invention] In summary, according to the invention, the average pressure in the exhaust passage when the exhaust brake is activated can be maintained high during high-speed engine operation, thereby increasing the exhaust braking force as much as possible. be able to.

[Brief explanation of the drawing]

Fig. 1 is a schematic configuration diagram showing a preferred embodiment of the exhaust brake device according to the present invention, Fig. 2 is a diagram explaining the ignition order of the engine shown in Fig. 1, and Fig. 3 is a diagram showing exhaust gas from each cylinder. Fig. 4 is an estimation graph of the waveform of the reflected wave of the pressure wave of Fig. 3; Fig. 5 is an estimation graph of the pressure waveform of the exhaust pulsation occurring at the exhaust port of the #1 cylinder; Figure 6 is a graph of the estimated pressure waveform of exhaust pulsation occurring at the exhaust port of the #7 cylinder, Figure 7 is a graph comparing the estimated waveform of exhaust pulsation in a V-type 8-cylinder engine with the measured waveform, and Figure 8 is A schematic configuration diagram showing a modification of the present invention, Fig. 9 is a graph of experimentally measured values showing the difference in exhaust braking force between a conventional exhaust brake device and an exhaust brake device according to the present invention, and Fig. 10 shows a collection of divided manifolds. A graph of experimentally measured values showing the difference in exhaust braking force when the exhaust brakes are assembled and when they are not collected. Figure 11 is an experimental measurement showing the difference in exhaust braking force that occurs when the length of the straight pipe part of the split manifold is changed. Figure 12 is a graph of experimentally measured values showing the influence of exhaust braking force due to differences in the shapes of the collecting pipes of split manifolds.
FIG. 3 is a schematic configuration diagram of a conventional exhaust brake device, and FIG. 14 is a diagram illustrating the operating principle of the exhaust brake. In the figure, #1, #2...#10 are cylinders, #1a,
#3a...#9a is an exhaust port, 15 is an exhaust passage, 17 is an exhaust manifold, 17a, 17b...
...17e is split manifold, 18#1, 18
#3...18 #9 is a branch pipe, 19a and 19b are straight pipes, 20 is a collecting pipe, and 21 is an exhaust brake valve.

Claims (1)

[Scope of utility model registration request] Each exhaust port of a multi-cylinder internal combustion engine is collected by a divided manifold for each cylinder where exhaust interference is difficult to occur, each divided manifold is further collected by a collecting pipe via a straight pipe section, and a pipe is opened and closed downstream of the collecting pipe. In an exhaust brake device equipped with an exhaust brake valve, the length of the straight pipe part of the split manifold is the length of the exhaust passage between the exhaust brake valve and the exhaust valve of each cylinder when the exhaust brake valve is closed during high-speed engine rotation. 1. An exhaust brake device characterized in that a valley portion of low gas pressure of exhaust pulsation that occurs pulsatingly within the exhaust brake device is formed to a length that acts on an exhaust valve of a cylinder during an intake stroke.