EP1242741A1

EP1242741A1 - Dual-stage, plunger-type piston compressor with minimal vibration

Info

Publication number: EP1242741A1
Application number: EP00991983A
Authority: EP
Inventors: Frank Meyer; Michael Hartl; Stefan Schneider
Original assignee: Knorr Bremse Systeme fuer Schienenfahrzeuge GmbH
Current assignee: Knorr Bremse Systeme fuer Schienenfahrzeuge GmbH
Priority date: 1999-12-21
Filing date: 2000-12-20
Publication date: 2002-09-25
Anticipated expiration: 2020-12-20
Also published as: EP1242741B1; US20030108435A1; KR20020065595A; DE50011039D1; ATE302906T1; US6776587B2; JP2003519740A; AU3726601A; KR100726202B1; JP4773022B2; CN1189658C; DE19961646C1; HK1054776A1; ES2248168T3; WO2001046585A9; CN1413292A; WO2001046585A1

Abstract

The invention relates to a piston arrangement for a dual-stage piston compressor, comprising a cranksbaft and several cylinders which house the operating piston. Said arrangement allows two or more low-pressure stages and at least one high-pressure stage to be formed. The invention is characterized in that the two or more low-pressure cylinders are arranged in relation to the high-pressure stage in such a way that said two or more low-pressure cylinders are in phase or are offset by less than ±15° and compress in a position which is offset by 180°±20, in relation to one or more high-pressure cylinders.

Description

Low-vibration, two-stage plunger compressor

The invention relates to a piston arrangement for a two-stage piston compressor with a crankshaft, a plurality of cylinders with pistons running therein, two or more low-pressure stages and at least one high-pressure stage being formed, and a piston compressor for rail vehicles with such a piston arrangement.

The arrangement according to DE-PS 765 994 is characterized in that the cylinders and the crank radii are designed in such a way that the inertial forces balance each other out as well as possible. Gas forces as vibration-inducing components are not mentioned. An assignment of the individual cylinders to a respective compression stage is not made in the document.

A piston compressor with a crankshaft, several cylinders and pistons running therein has become known, for example, from DE-PS 765 994.

Lightweight construction is increasingly being used in rail vehicle construction. Modern lightweight car body structures, for example made of extruded aluminum profiles or support structures made of thin sheet metal, often have natural frequencies close to the speed of the compressor of the air supply system. The use of piston compressors is often not possible in such constructions, since the permissible structure-borne noise level is often exceeded.

This is due to the fact that, due to their design, piston machines generate mass forces and moments caused by the oscillating masses on the crank mechanism and moments from the gas forces. A very non-uniform torque occurs particularly in the two-stage piston compressors frequently used in the rail vehicle sector. As the analysis of a typical load torque of such a compressor shows, the major part of the load torque corresponds to the rotational frequency of the piston engine, which is often in the range from 20 to 30 Hz. These frequencies are in turn for those in the passenger compartment People can be felt very well, because, for example, the natural frequency of legs with extended knees can be around 20 Hz.

The load torque of a piston compressor described above generates an excitation torque around the compressor axis of rotation in conjunction with the motor. The moment of inertia of a conventional piston compressor unit is much lower around the axis of rotation of the compressor than around other axes. Since the transmission mode of an elastic bearing around the longitudinal axis of the compressor is usually closer to the rotational frequency than, for example, the vertical mode, which plays a greater role for the transmission of inertial forces, this torsional vibration is usually not isolated as well as other excitation components.

The object of the invention is to provide a piston compressor machine which avoids the disadvantages described above.

According to the invention, the problem is solved by a drastic reduction in the proportion of the first order in the load moment predominantly resulting from the gas forces, in that two or more low-pressure stages overlap in phase and act approximately 180 offset from the high-pressure stage due to an unusual piston arrangement. This is structurally achieved in that in a piston arrangement for a two-stage piston compressor with a crankshaft a plurality of cylinders with pistons running therein, two or more low-pressure cylinders and at least one high-pressure cylinder being formed, the two or three or more low-pressure cylinders are arranged in this way with respect to the high-pressure cylinders are that the two or more low pressure cylinders compress in phase or offset by less than ± 15 and offset by 180 ± 20 to form one or more high pressure cylinders.

The inventors have recognized that the phase shift of all low-pressure cylinders to one or more high-pressure cylinders also results in a drastic reduction in the first order resulting from the torque diagram, and thus a drastic one Reduction of the vibrational moment around the compressor axis of rotation is achieved.

In a first embodiment of the invention, the piston arrangement is an oil-lubricated piston arrangement.

However, it is particularly preferred if the piston arrangement is an "oil-free", dry-running piston arrangement. In a special embodiment of the invention, the piston arrangement is designed as a 3-cylinder arrangement with two low-pressure and one high-pressure cylinders, a high-pressure cylinder being opposite another low-pressure cylinder. Such an arrangement is particularly space-saving. Of course, 4,5 or 6-cylinder arrangements which make use of the teaching according to the invention would also be conceivable.

According to an advantageous embodiment, the pressure peaks in the torque diagram can be significantly reduced by using heavy pistons, since an increased kinetic energy of the piston is converted into compression work. In particular, the pistons of the cylinders should have such a high mass that the pressure peaks in the tangential force diagram are reduced, the mass forces in the pressure peak in the tangential force diagram in the

Piston compressor typical speed range from 1000 1 / min to 2000 1 / min, in particular 1500 1 / min are greater than 15% of the gas forces in the pressure peak. In the present application, the tangential force diagram is understood to mean the torque curve / crank radius.

In an advantageous embodiment it is provided that, for example in a 3-cylinder arrangement, the masses of the low-pressure cylinder lying on the side of the high-pressure cylinder, namely the piston and / or connecting rod mass, are selected such that they oppose the low-pressure piston and the high-pressure piston. balance the two that sit on the same crankshaft crank. The compensation can take place here both on the piston and on the connecting rod. Due to the Increasing the piston mass as a result of the mass balance reduces the bearing load on the connecting rod.

In addition to mass balancing with the help of additional masses, it is also possible to balance the oscillating mass with the aid of idling pistons that move along. In the present application, a blind piston is understood to mean a piston that does no compression work.

The pistons are advantageously arranged such that the low-pressure pistons suck in phase over the crankcase, the two low-pressure stages diving into the crankcase pushing the air into the compression chamber during the suction process. This reduces the suction vacuum in the low pressure stage and improves the filling. In a particularly advantageous embodiment, this effect is reinforced by the use of a check valve on the inlet connection from the air filter housing to the crankcase. The arrangement of a check valve improves the efficiency, in particular of a dry-running piston arrangement.

In addition to the piston arrangement, the invention also provides a piston compressor, in particular for rail vehicles, comprising such a piston arrangement, which advantageously comprises an electric motor drive. The piston arrangement can also be used in compressed air generation systems in the industrial sector.

The invention will be explained below with reference to the drawings.

Show it:

Figure 1 shows the tangential force curve of a conventional two-stage piston compressor in a boxer arrangement, as it has become known, for example, from DUBBEL, paperback for mechanical engineering 15th edition or 18th edition, pages P 32 - P 33 Figure 2 shows the tangential force curve of a piston compressor according to the invention

Figure 3 shows a piston compressor according to the invention in section.

Figures 4a-4d possible embodiments of piston assemblies according to the invention designed as a box compressor

Figures 5a-5b embodiment of piston assemblies according to the

Invention carried out as a series machine

Figure 6 embodiment of a piston assembly with a blind piston

Figure 7 amplitudes of the compressor vibration in the vertical direction for an embodiment according to the prior art and the invention

FIG. 1 shows the tangential force diagram of a piston arrangement as known from the prior art, for example DUBBEL, paperback for mechanical engineering 15th edition or 18th edition pages P 32 - P 33. The x-axis designates the angle of rotation in degrees, the y-axis the applied torque. With 1 the moment from the gas forces is designated, with 3 the total moment from mass and gas forces and with 5 the moment from the mass forces.

The Fourier analysis of the load torque shown in FIG. 1 from the mass and gas forces of a compressor according to the prior art can be broken down into the following components:

1st order: 40 Nm

2nd order: 20 Nm 3rd order: 7 Nm

The majority of the load torque corresponds to the rotational frequency of the piston machine, which is often 20, 25 or 30 Hz. These frequencies are clearly noticeable to people, for example in the passenger compartment of a rail vehicle. The natural frequency of legs with extended knees can be approx. 20 Hz.

The load torque of a piston compressor, in cooperation with the motor, generates an excitation torque around the longitudinal axis of the compressor, the moment of inertia of a conventional piston compressor unit around the longitudinal axis of the compressor being substantially lower than around other axes. The transmission mode of an elastic bearing around the longitudinal axis of the compressor is generally closer to the rotational frequency than, for example, the vertical mode, which plays a greater role for the transmission of inertial forces. This torsional vibration is usually not isolated as well as other excitation components.

According to the invention, the vibration problem of conventional piston compressors is solved by a drastic reduction in the proportion of the first order in the load moment mainly resulting from the gas forces. This reduction in the first order can be achieved by means of a piston arrangement in which two or more low-pressure stages overlap in phase and act about 180 degrees apart from the high-pressure stage.

The tangential force diagram of such an arrangement is shown in FIG. 2. As in FIG. 1, reference number 1 denotes the moment from gas forces, 3 the moment from mass and gas forces and 5 the moment from mass forces.

The Fourier analysis of the curve according to FIG. 2 leads to the following result:

1st order: 19 Nm

2nd order: 28 Nm 3rd order: 7 Nm

The proportion of the first order is drastically reduced, which is reflected in a reduced vibration excitation around the longitudinal axis of the compressor. The undesirable vibrations in the passenger compartment can be considerably reduced or almost completely avoided.

FIG. 3 shows an example of a piston compressor with a piston arrangement according to the invention. The embodiment shown in FIG. 3 is, without limitation, a 3-cylinder boxer arrangement with two low-pressure cylinders 20, 22 which form the low-pressure stage and a high-pressure cylinder 24 which is arranged upstream of one of the low-pressure stages.

The pistons 40, 42, 44 of the three cylinders are mounted on a common crankshaft via connecting rods 32 with the help of ball or roller bearings 34.

To cool the arrangement, a fan wheel 36 is provided on the front side of the crankshaft 30, which ensures air cooling of the housing 38, in which the two low-pressure stages and the high-pressure stage are arranged, with the crankshaft 30 rotating.

In the position shown in FIG. 3, the pistons 40, 42 of the low-pressure cylinders are in the uppermost position. The high pressure piston 44 is located at the upper end of the cylinder. If the crankshaft 30 is moved, the two pistons 40, 42 of the low-pressure cylinders move in phase and are offset by 180 with respect to the piston 44 of the high-pressure stage.

The embodiment shown in FIG. 3 is an oil-free piston compressor with intake air routing through the crankcase. The individual pistons 40, 42, 44 are sealed off from the cylinder by means of sealing elements 50. The crankshaft is driven 30 with the aid of an electric motor 60.

The mode of operation of the piston compressor shown in FIG. 3 will be described in more detail below:

During the compression process in the low-pressure stages, the large low-pressure pistons 40, 42 which emerge from the crankcase 38 in phase increase the air volume in the crankcase 38. Air is sucked into the crankcase. When the air is sucked into the compression chamber, the low-pressure pistons 40, 42 dip into the crankcase 38. The volume in the crankcase 38 decreases at the moment that air is sucked out of the crankcase 38 into the compression chamber of the low-pressure stages, i.e. the piston underside of the low-pressure pistons 40, 42 pushes air out of the crankcase 38 into the compression chambers of the low-pressure stages. As a result, the suction vacuum in the low pressure stages is reduced compared to the designs according to the prior art. This effect can be supported when a check valve is used at the inlet connection of the air filter housing to the crankcase 38, the efficiency in particular being improved.

Another advantage of the piston compressor according to the invention is as follows:

Due to the strongly fluctuating load torque of the piston compressor, rotational irregularity is generated. This is amplified by the electric motor 60, because the motor 60 reacts out of phase to the load peak, specifically when the torque requirement of the compressor is low. The resulting speed fluctuation over one revolution can be, for example, ± 14% in the case of piston compressors according to the prior art. Until now, this effect could only be reduced by using large flywheels, which was not desirable for reasons of weight. Furthermore, the electric motor 60 has a significantly increased current consumption and a drastic reduction in the power factor - to to 0.6 - and must therefore be oversized in embodiments according to the prior art. This effect is reduced by the strong reduction of the first order in the load moment according to the invention. The rotational uniformity is reduced, it is reduced according to the invention from, for example, 0.15 to 0.08. The current consumption of the motor drops. The power factor increases considerably, for example from Lf = 0.7 to 0.8 in an arrangement according to the invention.

In a further developed embodiment, the use of heavy pistons can significantly reduce the pressure peaks in the torque diagram, since an increased kinetic energy of the piston is converted into compression work. It is particularly preferred if the pistons of the cylinders have such a high mass that the pressure peaks in the tangential force diagram are reduced, the mass forces in the pressure peak entering the tangential force diagram in the speed range between 1000 1 / min and 2000 1 / min greater than 15 % of the gas forces are in the pressure peak.

In this way it is possible to further reduce the excitation moments in all orders.

In order to achieve a mass balance of the oscillating and rotating masses, the masses of the low-pressure cylinder lying on the side of the high-pressure cylinder are selected so that they balance the opposite low-pressure piston and the high-pressure piston. The compensation can take place both on the piston and on the connecting rod. Increasing the piston mass as a result of the mass balance reduces the bearing load on the connecting rod. This is favorable for the load on the piston pin bearing of the low-pressure stage lying on the side of the high-pressure cylinder, because it is cooled more poorly by the adjacent high-pressure stage. FIGS. 4a-4d show arrangements according to the invention with opposing cylinders. FIG. 4a shows a 3-cylinder arrangement, as previously described in detail. FIG. 4b is a 6-cylinder arrangement, in FIG. 4c a 4-cylinder arrangement and in FIG. 4d a 5-cylinder arrangement according to the invention. The high-pressure pistons are given the reference numbers 44, 46 and the low-pressure pistons are given the reference numbers 40, 41, 42, 43. The high-pressure cylinders with the reference numbers 24, 26 and the low-pressure cylinders with the reference numbers 20, 21, 22, 23. In addition to 180 V piston arrangements, in-line machines are also conceivable.

Figure 5a shows an inventive 4-cylinder in-line machine, Figure 5b shows a 3-cylinder in-line machine.

FIG. 6 shows a 3-cylinder in-line engine with an idling piston 50, which does not perform any compression work and only serves as a mass balance. As in FIGS. 4a-4d, the high-pressure pistons are identified by 44 and the low-pressure pistons by 40, 42, the high-pressure cylinders by 24 and the low-pressure cylinders by 20, 22.

With the invention, therefore, for the first time, a piston arrangement and a piston compressor are made available, with which the undesired vibrations of the first order, such as result from the compression forces in piston compressors according to the prior art, can be reduced.

This can be seen particularly well in FIGS. 7a-7c. A compressor with two low-pressure cylinders 20, 22 and a high-pressure cylinder 24 according to the invention is shown schematically in FIG. 7a. Furthermore, four possible suspensions 70, 72, 74, 76 are shown, for example, on a rail vehicle. The cylinders lie in the x-y plane, the z-axis is perpendicular to the cylinder axis in the direction of the suspensions 70, 72, 74, 76.

FIG. 7b shows the time course of the 1st order compressor vibration in the z direction in a compressor according to the prior art. Figure 7c shows the Temporal course of the compressor vibration 1st order in the z direction in a compressor according to the invention. As can be seen from the comparison of the amplitudes of the vibrations in FIGS. 7b and 7c, the amplitude of the vibration of the compressor according to the invention is at least halved compared to the prior art. In a particularly preferred embodiment of the invention, the amplitude of a compressor according to the invention is only one third of the amplitude of a compressor according to the prior art.

LIST OF REFERENCE NUMBERS

1 moment from gas forces

3 moment from mass and gas forces

5 moment from mass forces

20, 22, 23 low pressure cylinders

24, 26, 28 high pressure cylinders

30 crankshaft

32 connecting rod

34 ball bearings

36 fan wheel

38 housing

40, 42, 43 low pressure pistons

44, 46, 48 high pressure pistons

49 sealing element

50 blind pistons

60 electric motor

70, 72, 74,

76 suspensions

Claims

claims

1. Piston arrangement for a two-stage piston compressor with

1.1 a crankshaft (30) 1.2 a plurality of cylinders (20, 22, 24) with pistons (40, 42, 43, 44, 46, 48) running therein, wherein

1.3 two or more low pressure cylinders and at least one high pressure cylinder is formed, characterized in that the two or three or more low pressure cylinders (20, 22, 23) are arranged with respect to the high pressure stage in such a way that the two or more

Compress low-pressure cylinders (20, 22, 23) in phase or offset by less than ± 15 and offset by 180 ± 20 to one or more high-pressure cylinders (24, 26, 28).

2. Piston arrangement according to claim 1, characterized in that the piston arrangement is an oil-lubricated piston arrangement.

3. Piston arrangement according to claim 1, characterized in that the piston arrangement is a dry-running piston arrangement.

Piston arrangement according to one of claims 1 to 3, characterized in that the piston arrangement as a 6-cylinder, 5-cylinder,

4-cylinder or 3-cylinder arrangement with two or more low-pressure cylinders (40, 42, 43) and one or more high-pressure cylinders (44, 46, 48).

5. Piston arrangement according to one of claims 1 to 4, characterized in that the piston arrangement is designed as a 3-cylinder arrangement which comprises two low (20, 22) and a high pressure cylinder (24), one high (24) and a low pressure cylinder (22) is opposite another low pressure cylinder (20).

6. Piston arrangement according to one of claims 1 to 5, characterized in that the pistons (40, 42, 43, 44, 46, 48) of the cylinders (20, 22, 23, 24, 26, 28) have such a high mass that the pressure peaks in the tangential force diagram are reduced, the mass forces entering the tangential force diagram in the pressure peak in the speed range between 1500 1 / min and 2000 1 / min being greater than 15% of the gas forces in the pressure peak.

7. Piston arrangement according to one of claims 1 to 6, characterized in that a mass balance of the oscillating masses (piston, connecting rod) with

Additional masses are carried out on at least one of the pistons (20, 22, 23, 24, 26, 28) and / or on a connecting rod.

8. Piston arrangement according to one of claims 1 to 7, characterized in that the mass compensation of the oscillating mass is carried out with the aid of idling pistons (50) which move along.

9. Piston arrangement according to one of claims 1 to 8, characterized in that the mass balance takes place with the aid of additional masses on the crankshaft (30).

10. Piston arrangement according to one of claims 1 to 9, characterized in that the piston arrangement is constructed such that the pistons of the low-pressure stage suck in phase through the crankcase (38), the two low-pressure pistons (40) immersed in the crankcase (38) during the suction process. 42, 43) force the air into the compression chamber.

11. Piston arrangement according to one of claims 1 to 10, characterized in that a check valve is arranged at the inlet opening to the crankcase (38).

12. Piston compressor, in particular for rail vehicles, characterized in that the piston arrangement comprises a piston arrangement according to one of claims 1 to 11.

13. Piston compressor according to claim 12, characterized in that the piston compressor comprises an electric motor drive (60).