U.S. patent application number 10/483019 was filed with the patent office on 2004-12-02 for compressor driveable by an electric motor.
Invention is credited to Jaisle, Jens-Wolf.
Application Number | 20040241018 10/483019 |
Document ID | / |
Family ID | 8164492 |
Filed Date | 2004-12-02 |
United States Patent
Application |
20040241018 |
Kind Code |
A1 |
Jaisle, Jens-Wolf |
December 2, 2004 |
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 w.sub.1 lies above the maximal operationally
occurring rotation speed n.sub.max.
Inventors: |
Jaisle, Jens-Wolf;
(Stuttgart, DE) |
Correspondence
Address: |
Stephan A Pendorf
Pendorf & Cutliff
P O Box 20445
Tampa
FL
33622-0445
US
|
Family ID: |
8164492 |
Appl. No.: |
10/483019 |
Filed: |
July 9, 2004 |
PCT Filed: |
July 6, 2001 |
PCT NO: |
PCT/EP01/07790 |
Current U.S.
Class: |
417/423.1 ;
417/423.12 |
Current CPC
Class: |
F04D 29/668 20130101;
F04D 25/06 20130101 |
Class at
Publication: |
417/423.1 ;
417/423.12 |
International
Class: |
F04B 017/00; F04B
035/04 |
Claims
1-6. (cancelled)
7. An 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, 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 W.sub.1 lies at least 25%, preferably 50%,
above the maximal occurring operational rotation speed
n.sub.max.
8. A compressor according to claim 7, wherein for raising the
fundamental frequency W.sub.1 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.
9. A compressor according to claim 8, wherein the bearing segments
(26) which are associated with the bearings (46) exhibit larger
dimensioned diameter.
10. A compressor according to claim 9, wherein the shoulder
segments (24) adjacent to the bearing segments (26) exhibit larger
dimensioned diameters.
11. A compressor according to claim 9, wherein 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).
12. A compressor according to claim 7, wherein the bearings (46)
are roller bearings.
Description
[0001] The invention concerns a compressor driveable by an electric
motor according to the precharacterizing portion of claim 1.
[0002] 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.
[0003] 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.
[0004] 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.
[0005] 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.
[0006] 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.
[0007] This problem is solved in a generic compressor by the
characterizing features of claim 1.
[0008] Advantageous embodiments of the invention are set forth in
the characteristics of the dependent claims.
[0009] 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.
[0010] 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 U.S. Pat. No.
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.
[0011] 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.
[0012] 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.
[0013] 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.
[0014] 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.
[0015] 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.
[0016] A further, additional option for raising the fundamental
frequency is comprised in designing the bearings arrangement of the
rotor shaft to be particularly stiff.
[0017] 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.
[0018] 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.
[0019] The invention will now be described in greater detail on the
basis of the embodiment schematically represented in the figures.
There is shown
[0020] FIG. 1 compressor in axial view;
[0021] FIG. 2 rotor shaft;
[0022] FIG. 3 deformation condition for a first rotor geometry;
[0023] FIG. 4 deformation condition for a second rotor
geometry.
[0024] The basic construction of a compressor can be seen from FIG.
1.
[0025] 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.
[0026] 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.
[0027] 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.
[0028] 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.
[0029] 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.
[0030] 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 w.sub.1, lies above the
maximal operating rotational speed n.sub.max. 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.
[0031] 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.
[0032] 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.
[0033] 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.
[0034] Further, an increase in the first critical bending
fundamental frequency w.sub.1 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.
[0035] A further possibility for raising of the first critical
bending fundamental frequency w.sub.1, 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.
[0036] 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.sub.1, is displaced to clearly
above the maximal occurring operational rotational speed n.sub.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 w.sub.1 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.
[0037] The success obtainable with the above described design
concept can particularly be seen by comparison of FIG. 3 and 4.
[0038] 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.
[0039] 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 w.sub.1, 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.sub.max.
Reference Number List
[0040] 1 Compressor
[0041] 10 Compressor wheel
[0042] 12 Securing nut
[0043] 14 Housing segment
[0044] 16 Flow channel
[0045] 18 Flow channel
[0046] 20 Rotor shaft
[0047] 22 Central segment
[0048] 24 Shoulder segment
[0049] 26 Bearing segment
[0050] 28 Seal segment
[0051] 30 Receiving segment
[0052] 32 Threaded segment
[0053] 42 Sealing disk
[0054] 44 Piston ring
[0055] 46 Roller bearing
[0056] 50 Stator
[0057] 52 Housing part
[0058] 54 Interstitial space
[0059] 56 Housing part
[0060] 58 Housing part
[0061] 60 Housing part
[0062] w.sub.1 first bend critical fundamental frequency
[0063] n.sub.max maximal operational rotational speed
* * * * *