COMPRESSOR DRIVEABLE BY AN ELECTRIC MOTOR

Abstract: Electric motor driveable compressor (1), in which the electric motor includes a rotor with rotor shaft (20), which acts in cooperation with a stator (50) provided in the motor housing (52, 56, 58, 60), the rotor shaft (20) being mounted on the motor housing (56, 58, 60) via at least two bearings (46) and connected fixed against rotation with the compressor wheel (10) at an axial receiving segment (30), wherein the rotating components, comprised essentially of the rotor with rotor shaft (20) and compressor wheel (10), are so designed, that their first critical bending fundamental frequency w1 lies above the maximal operationally occurring rotation speed nmax.

Description: COMPRESSOR DRIVEABLE BY AN ELECTRIC MOTOR Description The invention concerns a compressor driveable by an electric motor according to the precharacterizing portion of Claim 1. This type of compressor is driven by an electric motor, which can be constructed for example as an asynchronous or induction motor. It is comprised essentially of a rotor with a rotor shaft, which works in cooperation with a corresponding stator provided in the motor housing. The rotor shaft is supported in the motor housing via two journals or bearings, and exhibits an axial mounting or receiving segment, onto which a compressor wheel is provided fixed against rotation. Energizing the motor causes the rotor shaft to rotate, which in turn directly drives the compressor wheel. The axial segments which receive the mounting rings as well as in many cases the necessary shaft seal rings are designed to be as small as possible with regard to their diameter. This applies in like manner to the shoulder segments immediately adjacent to the mounting points, on which the mounts, for example the inner mount rings, are brought to bear. The diameters are so dimensioned, that they are capable of standing up to the demands occurring during the projected product life. Modern design processes, which utilize intelligent or cascade simulation programs, make possible a relatively reliable prediction of the component behavior, so that the diameter of the axial segment of interest for a given rotor shaft can be constructed or designed to be relatively small . Beyond this, small diameters are considered desirable for rotating construction components, since this makes possible a reduction of the rotation inertia. This results in an optimal responsiveness during charges in rotational speed, in particular during accelerating or decelerating the compressor. Although this type of compressor has fundamentally proven itself in practice, it however exhibits certain disadvantages. It is particularly noted, that especially during acceleration to the operating rotational speed, vibrations occur, which lead not only to undesirable noise emissions, but rather also reduce the life expectancy, caused for example by damage to the bearings. The present invention is thus concerned with the problem of improving the compressor of the above-described type in such a manner that the mentioned disadvantages no longer occur. In particular, the running quietness and noise emissions should be improved and the life expectancy should be increased. This problem is solved in a generic compressor by the characterizing features of Claim 1. Advantageous embodiments of the invention are set forth in the characteristics of the dependent claims. The invention is based upon recognition of the fact that in the application of design criteria until now the rotating group of components, comprised essentially of the rotor with the rotor shaft and the compressor wheel, have a first critical bending fundamental frequency or resonance frequency, which lies within the operational rotational speed spectrum. During acceleration of the compressor this fundamental frequency must be passed through, at which time the vibrations and noise emissions are triggered. Measures are known from other technical fields, for example, the driving of hard disks for a computer, which are directed to a targeted dampening of oscillations which are emitted by components rotating at a high rotational speed. For example, in US 6,140,790 a complex governing process is described for dampening vibrations in a rotating system. An application to the present case of electric motor driven compressors might appear possible, but requires however a complex supplemental design and control means. In comparison to this, the present invention makes it possible to avoid the vibrations occurring in the conventionally designed, electric motor driven compressors, in simple manner thereby, that the first critical bending harmonic frequency is raised to a value which lies above the maximal rotational speed occurring during operation. In this manner substantial improvements in the operational behavior, in particular with respect to running quietness, noise emissions and operational life, can be achieved. This is achieved without the usual conventional supplemental measures, such as for example active or passive dampening, so that these advantages can be achieved practically without supplemental construction complexity or investment . Preferably, the raising in the harmonic frequency is achieved by modification of the rotor shaft. By targeted modification in the manner of an enlarging of the diameter of the axial segments, it becomes possible to economically achieve the desired frequency displacement, since the rotor shaft is designed as a rotating component and accordingly can be modified without problem with respect to the diameter of the axial segments. Those axial segments which receive the bearings have proven themselves as optimal starting point for modifications of this type. These bearing segments are designed or constructed with larger dimensioned diameter. Essentially, the inner diameter of the bearing provides a limit, which depends upon the highest permissible rotational speed of the bearings. A raising of the fundamental frequency is also produced as a consequence of increasing the dimensions of the shoulder segments adjacent to the mounting segments. The shoulder segments respectively serve for the abutment of the inner ring of the roller bearing, so that any enlargement of the diameter of the bearing segment necessarily also produces a corresponding enlargement of the diameter of the adjacent shoulder segment. Further, the fundamental frequency can be increased by enlargement of the axial segments which receive the seal disks. It is understood that the diameter of this bearing segment may at most be equal to the diameter of the adjacent bearing segments, since otherwise an introduction of the inner bearing rings is no longer possible. A further, additional option for raising the fundamental frequency is comprised in designing the bearings arrangement of the rotor shaft to be particularly stiff. With . the aid of the preceding described options it becomes possible to vary the fundamental frequency within wide ranges and in particular to displace it to an area, which is a sure distance from the maximal occurring operational rotational speed. The safe distance of the fundamental harmonic to the highest operational rotational speed is a function of, among other things, how much the manufacturing and friction dependent oscillations must be taken into consideration. It has been found advantageous to displace the fundamental frequency by at least 25%, preferably however by 50% above the maximal rotational speed occurring during operation. The invention will now be described in greater detail on the basis of the embodiment schematically represented in the figures. There is shown Fig. 1 compressor in axial view; Fig. 2 rotor shaft; Fig. 3 deformation condition for a first rotor geometry; Fig. 4 deformation condition for a second rotor geometry. The basic construction of a compressor can be seen from Fig. 1. The compressor 1 includes a compressor wheel 10 which is driveable via rotor shaft 20. For this, the rotor shaft 20 exhibits a receiving segment 30, upon which the compressor wheel 10 is seated and is secured on the rotor shaft against rotation via a securing nut 12, which is screwed onto a threaded segment 32 of the rotor shaft 20. Air channels 16, 18 are formed by an appropriate design of the compressor housing segment 14. The rotor shaft 20 includes a central segment 22. A corresponding stator part 50 is provided. Interstitial spaces 54 serve as coolant water channels. In this manner an asynchronous motor is constructed, which serves for driving the compressor wheel 10. A motor control, not shown here, is provided in housing part 60, which closes the end side of the compressor 1. For mounting the rotor shaft 20, two roller bearings 46 are provided. The rotor shaft 20 includes bearing segments 26, which are designed as bearing seats with close tolerances. Likewise, the housing parts 56, 58 in the area of the roller bearings 46 are manufactured with close tolerances, whereby a defined fit results. Adjacent to the bearing segments 26 shoulder segments 24 are provided which transition to the central segment 22, and against which the roller bearings 46 respectively are supported axially. The rotor shaft 20 exhibits a seal segment 28 between the receiving segment 30 and the bearing segment 26, which seal segment carries a seal disk 46 with a piston ring 44. Thereby a seal is formed between the air channel and the area of the asynchronous motor. The invention is manifested therein, that the rotating construction components, comprised essentially of the rotor shaft 20 and compressor wheel 10, are so designed, that its first critical bending fundamental frequency wi lies above the maximal operating rotational speed n ma χ. In the stage of the numeric design there are however a series of peripheral conditions to be observed, so that the design of the rotor shaft 20, as is also shown in Fig. 2, has significant meaning. With respect to the design of the compressor wheel 10 aero- thermodynamic preconditions are to be observed or maintained, which in connection with the material characteristics leave hardly any flexibility for a targeted influencing of the geometry of the compressor wheel 10. Thus the displacement of the first critical bending fundamental frequency is achieved by the targeted influencing of the geometry of the rotor shaft 20. In the modification of the rotor shaft 20 it is first to be observed that the maximal diameter in the area of the central segment 22 in general is predetermined, since it is dependent upon the geometry of the stator, in particular the stator part 50. These values are determined by the electrical power data of the asynchronous motor, and with respect to the required drive requirements are not variable. A first possibility for influencing the first critical bending fundamental frequency thus lies in the enlargement of the diameter in the area of the shoulder segment 24. The diameter enlargement is essentially subject to a limit with respect to the housing contour running in this area. Further, an increase in the first critical bending fundamental frequency wi is possible by an enlargement of the diameter in the area of the mounting or bearing segment 26. The maximal possible diameter is influenced by the design of the bearing and limited by the maximal possible internal diameter of the lower bearing 46. It must be observed, that the diameter of a mounting or bearing segment 26 and the shoulder segment 24 must, at least in the transition area, be coordinated or adapted with respect to each other, so that the axial supporting of the roller bearing 46 can satisfy design requirements . A further possibility for raising of the first critical bending fundamental frequency wi can finally also occur by a corresponding enlarging of the diameter in the area of the seal segment 28. The diameter may however maximally reach the value of the diameter in the area of the adjacent bearing segment 26, since during the assembly of the motor the roller bearing 46 must be slid axially over the seal segment 28 and onto the bearing segment 26. The application of the above criteria leads to a rotor shaft 20, which is compact in comparison to the conventionally designed rotor shafts. By the targeted enlargement of the diameter in the axial segments of interest it is accomplished that the critical bending fundamental frequency w is displaced to clearly above the maximal occurring operational rotational speed n max . This has the result, that the compressor 1, during acceleration to the operating speed, no longer passes through the first critical bending fundamental frequency Wi of the rotating construction components, and as a result the operating behavior with respect to running quietness, noise emission and product life is significantly improved. The success obtainable with the above described design concept can particularly be seen by comparison of Fig. 3 and 4. Fig. 3 shows a first construction component group, comprised of compressor wheel 10 and rotor shaft 20, which is symbolized by a network. The representation shows a so-called "screenshot" of a numeric simulation, in which the deformation corresponding to the first critical bending characteristic of the network is represented exaggerated. The repeated parameters show a frequency of 1210.71 hertz. The rotating construction component shown in Fig. 4 is, with respect to the compressor wheel 10, provided with the same design. The differences are concerned with the modified axial segment of the rotor shaft 20 in the above described mode and manner. The screenshot according to Fig. 4 shows that as a result of the comparatively small modification to the rotor shaft 20 the first critical bending fundamental frequency can be displaced to 2124.67 hertz. The simulation shows that it is possible, with comparatively small expense and complexity, to raise the first critical bending fundamental frequency wi by a factor of 2 and therewith to displace it to an area which, in the present case, lies above the maximal operational rotational speed n max . Reference Number List 1 Compressor 10 Compressor wheel 12 Securing nut 14 Housing segment 16 Flow channel 18 Flow channel 20 Rotor shaft 22 Central segment 24 Shoulder segment 26 Bearing segment 28 Seal segment 30 Receiving segment 32 Threaded segment 42 Sealing disk 44 Piston ring 46 Roller bearing 50 Stator 52 Housing part 54 Interstitial space 56 Housing part 58 Housing part 60 Housing part Wi first bend critical fundamental frequency n max maximal operational rotational speed

Claims: Patent Claims 1. Electric motor driveable compressor in which the electric motor includes a rotor with rotor shaft, which acts in cooperation with a stator provided in the motor housing, the rotor shaft is mounted in the motor housing via at least two bearings and is connected fixed against rotation with the compressor wheel at an axial receiving segment, thereby characterized, that the rotating components, comprised essentially of the rotor with rotor shaft (20) and compressor wheel (10) , are so designed, that their first critical bending fundamental frequency wi lies above the maximal operationally occurring rotation speed n max . 2. Compressor according to Claim 1, thereby characterized, that for raising the fundamental frequency Wi the rotor shaft (20) includes at least one axial segment (24, 26, 28), of which the cross sectional diameter is larger than the diameter dimensioned in accordance with conventional design criteria. 3. Compressor according to Claim 2, thereby characterized, that the bearing segments (26) which are associated with the bearings (46) exhibit larger dimensioned diameter. 4. Compressor according to Claim 3, thereby characterized, that the shoulder segments (24) adjacent to the bearing segments (26) exhibit larger dimensioned diameters. 5. Compressor according to Claim 3, thereby characterized, that the sealing segments (28) adjacent to the bearing segments (26) exhibit a diameter, which is equal to or slightly smaller than the diameter of the bearing segments (26) . 6. Compressor according to one of the preceding claims, thereby characterized, that the bearings (46) are roller bearings . 7. Compressor according to one of the preceding claims, thereby characterized, that the fundamental frequency i lies at least 25%, preferably 50%, above the maximal occurring operational rotational speed n max .

Inventors: JAISLE JENS-WOLF, JAISLE, JENS-WOLF
Applicants: BORG WARNER INC, JAISLE JENS-WOLF, BORGWARNER INC., JAISLE, JENS-WOLF
Application year: 2001
Main classification: F04D25
IPC codes: F04D25, F04D29, H02K7