U.S. patent application number 11/814738 was filed with the patent office on 2008-06-12 for axial piston compressor.
This patent application is currently assigned to VALEO COMPRESSOR EUROPE GMBH. Invention is credited to Henrick Brandes, Ullrich Hesse, Otfried Schwarzkopf, Oliver Tschismar.
Application Number | 20080138212 11/814738 |
Document ID | / |
Family ID | 36072227 |
Filed Date | 2008-06-12 |
United States Patent
Application |
20080138212 |
Kind Code |
A1 |
Schwarzkopf; Otfried ; et
al. |
June 12, 2008 |
Axial Piston Compressor
Abstract
Compressor, especially a compressor for the air-conditioning
system of a motor vehicle, having a housing (1) and, for drawing in
and compressing a coolant, a compressor unit arranged in the
housing (1) and driven by means of a drive shaft, the compressor
unit being regulated by means of the pressure (P.sub.C) prevailing
in a drive mechanism chamber substantially bounded by the housing
(1), there being an additional regulation device (17) and/or
control device for the inlet-gas-side coolant mass flow and/or the
inlet pressure and/or the inlet density.
Inventors: |
Schwarzkopf; Otfried;
(Kuerten, DE) ; Hesse; Ullrich; (Affalterbach,
DE) ; Brandes; Henrick; (Ludwigsburg, DE) ;
Tschismar; Oliver; (Weil im Schoenbuch, DE) |
Correspondence
Address: |
VOLPE AND KOENIG, P.C.
UNITED PLAZA, SUITE 1600, 30 SOUTH 17TH STREET
PHILADELPHIA
PA
19103
US
|
Assignee: |
VALEO COMPRESSOR EUROPE
GMBH
Hockenheim
DE
|
Family ID: |
36072227 |
Appl. No.: |
11/814738 |
Filed: |
January 25, 2006 |
PCT Filed: |
January 25, 2006 |
PCT NO: |
PCT/EP06/00646 |
371 Date: |
September 17, 2007 |
Current U.S.
Class: |
417/222.1 ;
417/270; 92/13 |
Current CPC
Class: |
F04B 2207/044 20130101;
F04B 27/1804 20130101; F04B 2205/171 20130101; F04B 2027/1877
20130101; F04B 49/225 20130101; F04B 2205/01 20130101; F04B
2027/1872 20130101 |
Class at
Publication: |
417/222.1 ;
92/13; 417/270 |
International
Class: |
F04B 27/18 20060101
F04B027/18; F04B 49/22 20060101 F04B049/22; F04B 1/12 20060101
F04B001/12; F04B 49/12 20060101 F04B049/12 |
Foreign Application Data
Date |
Code |
Application Number |
Jan 25, 2005 |
DE |
102005003494.2 |
Feb 21, 2005 |
DE |
102005007849.4 |
Claims
1. A compressor comprising a housing (1) and, for drawing in and
compressing a coolant, a compressor unit arranged in the housing
(1) and driven by a drive shaft, the compressor unit being
regulated by a pressure (P.sub.C) prevailing in a drive mechanism
chamber substantially bounded by the housing (1), and an additional
regulation device (17) and/or control device for an inlet-gas-side
coolant mass flow and/or an inlet pressure and/or an inlet
density.
2. Compressor according to claim 1, wherein the regulation device
(17) comprises a throttling location having an adjusting member
(13) comprising a throttling valve or a throttling flap, or
comprises a pressure reducer.
3. Compressor according to claim 2, wherein the adjusting member
(13) of the regulation device adjusts the coolant mass flow or,
that is, the inlet pressure in dependence on a speed of
rotation.
4. Compressor according to claim 2, wherein the throttling location
has an end-stop, associated with the adjusting member (13), for a
position of minimum flow cross-section such that even in the case
of very high speeds of rotation a predetermined minimum coolant
mass flow or inlet pressure is ensured.
5. Compressor according to claim 2, wherein the adjusting member is
an adjusting piston (13).
6. Compressor, especially according to claim 1, wherein the control
device comprises at least one inlet valve arranged on an inlet gas
side.
7. Compressor according to claim 6, wherein the inlet valve is a
pressure-controlled reed valve.
8. Compressor according to claim 6, wherein the inlet valve
comprises a valve plate (18) having a throttling through-bore (19)
and a tongue-like inlet blade (21).
9. Compressor according to claim 8, further comprising a cylinder
block and at least one piston which is arranged to move axially
back and forth in a corresponding at least one bore provided in the
cylinder block, the inlet valve comprises at least one inlet valve
so that a separate one of the inlet valves is associated with each
cylinder and corresponding inlet blades (21) are integrated in an
inlet blade plate.
10. Compressor according to claim 9, wherein an end of each
cylinder space which is associated with a respective one of the at
least one inlet valve has a radially extending annular extension,
which limits a stroke of the associated inlet blade (21) and which
is bevelled off or flattened off towards a fixing location of the
associated inlet blade.
11. Compressor according to claim 6, further comprising a cylinder
block and at least one piston, which is arranged to move axially
back and forth in respective bores provided in the cylinder block,
a ratio of piston diameter and piston stroke (D/s) is approximately
from 0.4 to 1.5.
12. Compressor according to claim 11, wherein a ratio of piston
diameter and a throttling through-bore in a valve plate (D/d) is
approximately from 1.5 to 5.
13. Compressor according to claim 12, wherein a ratio of the
throttling through-bore in the valve plate and a stroke of an inlet
blade (d/t) is approximately from 2.5 to 8.
14. Compressor according to claim 13, wherein a ratio of piston
stroke to the stroke of the inlet blade (s/t) is approximately from
10 to 30.
15. Compressor according to claim 6, wherein the control device is
defined by the geometry of the inlet valve.
16. Compressor according to claim 1, wherein the control device
comprises an orifice plate arranged on a inlet gas side.
17. Compressor according to claim 1, wherein the compressor unit
comprising pistons (4) moving axially back and forth in a cylinder
block (2) and, driving the pistons (4) and rotating together with
the drive shaft (6), a tilt plate (7), especially a swash plate or
wobble plate or a tilt ring, a deflection angle of the tilt plate
(7) is governed by interaction of, on the one hand, a pressure in a
drive mechanism chamber (8') substantially accommodating the tilt
plate and, on the other hand, the coolant mass flow on the inlet
side or the inlet pressure.
18. Compressor according to claim 1, wherein the regulation device
(17) is arranged to be actuated or controlled from outside the
compressor.
19. Compressor according to claim 1, wherein the regulation device
(17) and/or control device include(s) an inlet-gas-side oil
separator.
20. Compressor according to claim 1, wherein the regulation device
(17) is self-regulating in dependence on a difference in pressures
at, on the one hand, an outlet side or high-pressure side and, on
the other hand, an entry side or inlet-pressure side.
21. Compressor according to claim 17, wherein a moment distribution
of components of the compressor that are moved in rotation and in
translation is such that, in the case of an increase in a
compressor speed of rotation, the deflection angle of the tilt
plate (7) remains substantially the same or decreases.
22. Compressor according to claim 17, wherein the regulation device
and/or control device is/are arranged in an inlet gas channel (9)
extending mainly in a cylinder block (2).
23. Compressor according to claim 22, wherein the regulation device
(17) comprises a coolant mass flow measuring device and/or a
pressure sensor in the inlet gas channel (9) to one side or to each
side of the throttling location and/or the speed of rotation of the
compressor and/or of an engine driving the latter and/or a pressure
on an outlet side of the compressor.
24. Compressor according claim 23, wherein the adjusting member
(13) of the regulation device (17) acts against a force of a
resilient element.
25. Compressor according to claim 24, wherein the force of the
resilient element or biasing exerted by the resilient element on
the adjusting member (13) is adjustable by an adjusting screw
(15).
26. Compressor according to claim 25, wherein the adjusting member
(13) is arranged between the pressure-gas side and the inlet gas
side.
27. Compressor according to claim 17, wherein the regulation device
comprises a throttling location of constant cross-section.
28. Compressor according to claim 27, wherein an inlet line and/or
a connection between the compressor and an evaporator are part of
the regulation device.
29. Compressor according to claim 28, wherein across the regulation
device there is a pressure difference of approximately 1 bar at a
compressor speed of rotation of approximately 600 rpm and/or of
approximately 10 bar at approximately 8000 rpm.
30. Compressor according to claim 28, wherein the regulation device
or the throttling location comprises a tubular line having a
tubular cross-section of approximately from 8 to 10 mm.
31. Compressor according to claim 30, wherein CO.sub.2 is used as
coolant.
32. Compressor according to claim 17, wherein a moment due to
components of the compressor that are moved in rotation M.sub.SW is
of substantially equal magnitude to a moment M.sub.k,ges due to
components of the compressor that are moved in translation.
Description
[0001] The invention relates to an axial piston compressor,
especially to a compressor for the air-conditioning system of a
motor vehicle, in accordance with the preamble of claim 1.
[0002] From DE 197 49 727 A1 there is known an example of an axial
piston compressor of such a kind, having a housing and, arranged in
the housing and driven by way of a drive shaft, a compressor unit
for drawing in and compressing a coolant. The compressor unit is
regulated substantially by the pressure P.sub.C in a drive
mechanism chamber depending in the particular instance on the load
and/or speed of rotation of the compressor, an inlet pressure
P.sub.V1 and a high pressure P.sub.V2 that prevail on the inlet and
outlet sides, respectively, of the compressor also having an
influence on regulation of the compressor. Regulation takes place
by way of a change in the piston stroke of the compressor, the
piston stroke being governed by the deflection of a tilt plate from
a zero position.
[0003] In the case of such an axial piston compressor it is
additionally the case, however, that because of the force
conditions prevailing in the compressor, especially because of
centrifugal forces, the compressor has a tendency to up-regulate at
high speeds of rotation, that is to say the compressor has a
tendency towards a greater piston stroke and accordingly towards a
higher pressure on the high-pressure side.
[0004] DE 195 14 748 C2 explains the tilting moments which in the
case of, in principle, all compressors existing in the prior art of
the tilt plate form of construction act on the tilt plate and which
decisively contribute to the tilting or, that is to say,
inclination behaviour of the tilt plate. The compressor tilting
behaviour described in that publication can be regarded as being
exemplary of compressors of the tilt plate form of construction. In
general it is the moments given below which, in the centre of the
tilting movement of the tilt plate, have an influence on the
tilting of the tilt plate. The direction of the moment is given in
brackets, with (-) denoting down-regulation, that is to say
regulation in the direction of minimum stroke, and (+) denoting
up-regulation, that is to say regulation in the direction of
maximum stroke of the pistons. It is substantially the following
moments that play a part: [0005] moment due to gas forces in the
cylinder spaces (+) [0006] moment due to gas forces from the drive
mechanism chamber (-) [0007] moment due to a restoring spring (-)
or [0008] moment due to an advancing spring (+) [0009] moment due
to rotating masses (-), these including, for example, the tilt
plate (including a moment due to the position of the centre of
gravity, in which case--for example in accordance with DE 195 14
748 C2--this component can be positive, that is to say (+)) [0010]
moment due to masses moved in translation (+), which can include,
for example, pistons, sliding blocks or also an oscillating wobble
plate.
[0011] As can be seen from the above listing, the moment due to the
rotating masses, hereinafter referred to as M.sub.SW, generally has
a down-regulating action over an extensive tilt angle range. It is
only in the region of very small tilt angles, as a result of, for
example, an outwardly displaced location of centre of gravity
(Steiner component in the calculation of the moment of deviation
J.sub.yz) that an up-regulating moment can be produced in the case
of the tilt plate. DE 195 14 748 C2 also shows a plot of the moment
due to masses moved in translation, which as already explained has
an up-regulating action.
[0012] Also of interest is the sum of the moments, which in the
case of the subject-matter of DE 195 14 748 C2 is responsible for
up-regulating behaviour of the compressor over the entire tilt
angle range of the tilt plate, because the masses moved in
translation dominate the regulation behaviour in a wide tilt angle
range.
[0013] With reference to the prior art, especially according to DE
195 14 748 C2, it is disadvantageous therein that, when there is an
increase in the delivery volume due to an increase in the speed of
rotation, an additional increase in the delivery volume due to an
increase in the tilt angle of the tilt plate is added thereto. This
has to be compensated by appropriate interventions for the purpose
of regulation, which is onerous, reduces the efficiency of the
driving engine and accordingly increases the fuel consumption.
[0014] From EP 0 809 027 A1 there is known a compressor in which an
attempt is made so to compensate the delivery quantity of the
compressor by means of the dynamic behaviour of the compressor
drive mechanism that the delivered quantity of coolant or, that is
to say, the coolant mass flow can be kept constant. For regulating
the delivery quantity so that it is constant in the case of
changing speeds of rotation it is proposed that the restoring
torque of the wobble plate be utilised, the inclined position of
which is counteracted by the restoring torque by virtue of dynamic
forces on the co-rotating plate part.
[0015] Measures are known from DE 198 39 914 A1 as to how
regulation behaviour of such a kind--that is to say, at least
partial compensation of the delivery quantity--can be achieved. It
is proposed that the tilt plate component mass be so dimensioned in
relation to the masses moved in translation that the centrifugal
forces of the tilt plate influence the regulation behaviour of the
tilt plate. According to DE 198 39 914 A1 it is proposed that the
rotating mass of the tilt plate or, that is to say, of the tiltable
part of the tilt plate be greater than the total mass of all the
pistons so that the centrifugal forces occurring during rotation of
the tilt plate are sufficient to counteract the tilting movement of
the tilt plate in regulating manner deliberately and thereby to
influence, especially to reduce or to limit, the piston stroke and,
accordingly, the delivery quantity.
[0016] In DE 103 29 393 A1 belonging to the Applicant it is
furthermore explained why the component mass should not be the
preferred parameter for influencing as desired the regulation
behaviour of the drive mechanism resulting from changes in the
speed of rotation. According to DE 103 29 393 A1, the desired
regulation behaviour of the compressor is primarily achieved not by
means of the tilt plate component mass in relation to the masses
moved in translation but rather by taking into account the mass
moment of inertia of the tilt plate arrangement, which depends more
on its geometry than on its component mass. A central concept in
that Application is that, in the case of fluctuations or changes in
the speed of rotation, the moment due to masses moved in
translation is directly compensated, or even over-compensated, by
the moment due to rotating masses. In the case of new compressors
it is in fact desirable to reduce to a low level the frequency and
intensity of interventions for the purpose of regulation.
[0017] At this point reference should again be made to EP 0 809 027
A1, in which the objective is set of achieving regulation of the
delivery quantity so that it is constant. It can, however, be
simply demonstrated that appropriate down-regulation is not
possible solely by means of the advancing moment acting on the tilt
plate. The delivery volume is directly proportional to the speed of
rotation, which is to say, therefore, that when the speed of
rotation doubles the delivery volume also doubles. However, for the
tilt plate tilt moment, which is produced by the relevant moment of
deviation, the following equation applies:
M.sub.SW=J.sub.yz.omega..sup.2. Because it is the square of the
speed of rotation that influences the tilt moment, the aim of
regulating the delivery quantity so that it is constant cannot be
achieved solely by designing or dimensioning an appropriate tilt
plate.
[0018] The subject-matter of DE 198 39 914 A1 and of DE 103 29 393
A1 already discussed hereinbefore shows approaches for solutions to
and arrangements for the objective formulated in EP 0 809 027 A1.
In tendential terms it is possible in the case of fluctuations in
the speed of rotation, for example therefore in the case of an
increase in the speed of rotation at the compressor shaft, for the
tilt angle of the tilt plate either to increase (for example, in
accordance with DE 195 14 748), which is undesirable, or else in
accordance with DE 198 39 914, DE 103 29 393 or EP 0 809 027 A1 to
decrease, in which case it should be mentioned that even a decrease
or a decreasing tendency in the tilt angle or, that is to say,
deflection angle of the tilt plate is of only qualified
desirability. In any event, the desired objective is possible only
by means of a complicated arrangement which also has to be balanced
in highly precise manner especially so that the moment
distributions are taken into account. This results in compressors
according to the prior art being relatively expensive to
manufacture.
[0019] In the (unpublished) Patent Application DE 103 47 709
belonging to the Applicant, it is proposed that the active moments
due to the mass forces or, that is to say, the moments due to the
moments of deviation be so matched that the deflection angle of the
tilt plate does not change substantially in the case of changing
speeds of rotation. In the case of the subject-matter of DE 103 47
709 it was recognised that regulation behaviour of such a kind
represents optimal drive mechanism behaviour in order, by that
means, to be able to regulate the mass flow of a coolant compressor
in optimal manner.
[0020] FIG. 15 shows in diagrammatic form how a compressor 101 of
the described form of construction is regulated. Such a compressor
makes available, in operation, an inlet gas pressure level and also
a high-pressure level. The coolant circuit also has these same
pressure levels. A certain pressure modification or, that is to
say, pressure adjustment is accomplished by means of an expansion
element 103, which in turn reacts to changes in the operating state
of the compressor and, where appropriate, intervenes for the
purpose of regulation. In the compressor drive mechanism chamber
there is established, for example by means of regulatory valves of
the compressor, a pressure which is between the inlet gas pressure
level and the high-pressure level. The drive mechanism chamber
pressure intervenes in the force equilibrium or, that is to say,
moment equilibrium at the tilt plate in such a way that the tilt
angle of the tilt plate can be modified. If the drive mechanism
chamber pressure is set slightly above the inlet pressure, the tilt
plate is adjusted to the maximum tilt angle. If the drive mechanism
chamber is set substantially above the inlet pressure, the tilt
angle is adjusted to the minimum tilt angle. Regulation is
accomplished by means of the possible volume flows between the
individual chambers or pressure levels. Furthermore, reference
numeral 102 denotes a gas cooler/liquefier, reference numeral 104
denotes an evaporator and reference numeral 105 denotes a
regulation path. As an alternative to the regulation path shown by
the solid line between the gas cooler/liquefier 103 and the
pressure level P.sub.V1, wherein P.sub.V1 is selected as the
desired value, there is shown, by a broken line, a second
possibility for the regulation path 105, which has P.sub.V2 as the
desired value. A regulation path of such a kind is more usual
especially for the coolant CO.sub.2. The described model is shown
here in simplified form and is to be considered as being by way of
example.
[0021] Because, in operation of the compressor or, that is to say,
operation of the vehicle, the speed of rotation of the compressor
drive mechanism or, that is, of the vehicle engine changes almost
constantly, interventions for the purpose of regulation are
constantly required in the case of compressors according to the
prior art in order, for example, to ensure a constant delivery
output of the compressor or to maintain the desired value. Because
the drive mechanism chamber has a comparatively large volume,
regulation by means of appropriate modification of the drive
mechanism chamber pressure is sluggish and substantial overshoot
occurs. Consequently, a constant delivery output of the compressor
can be achieved only under difficult circumstances. Likewise, the
interventions for the purpose of regulation reduce the performance
of the vehicle and consume power of the driving engine.
[0022] Starting from the prior art discussed hereinbefore, the
objective of the present invention is to provide a compressor
(especially, but not exclusively, a tilt plate compressor having
variable piston stroke), for use in vehicle air-conditioning
systems, which has a speed of regulation that is substantially
improved compared to the prior art and which can keep the coolant
mass flow constant over wide speed of rotation ranges without
considerable loss of performance.
[0023] The objective is met in accordance with the invention by a
compressor having the features of patent claim 1, advantageous
developments and details of the invention being described in the
subordinate claims.
[0024] It is accordingly a fundamental point of the present
invention that a compressor having a compressor unit in accordance
with the preamble of patent claim 1 has an additional regulation
device and/or control device for the inlet-gas-side coolant mass
flow and/or the pressure on the inlet side and/or the inlet
density. A constructional measure of such a kind brings about, in
addition to throttling of the coolant mass flow, a down-regulating
effect (constituting the major part of the total effect), which is
produced by utilising the pressure difference--brought about by the
throttling--between the pressure side or, that is to say, the
pressure (P.sub.V2) prevailing on the outlet side and the inlet
side or, that is to say, the pressure level P.sub.V1* acting on the
pistons. For the purposes of illustration, a construction according
to the invention applied to an axial piston compressor will be
considered hereinbelow, although it should be noted that this must
not in any way be interpreted in limiting manner because a
construction according to the invention can likewise be applied to
a whole series of compressors of some other form of construction.
If the coolant flows into the cylinders of the axial piston
compressor at a relatively low pressure level or, that is, a
relatively low inlet density, the compressor has the tendency,
given constant pressure in the drive mechanism chamber, to
down-regulate, that is to say to reduce the piston stroke.
Accordingly, limitation of the coolant mass flow or, that is, of
the inlet pressure intervenes directly in the regulation of the
compressor. Expressed in other words, a modification of the volume
flow results for the main part in a change in the pressure
difference which regulates the compressor and, therefore, in a
predetermined regulation characteristic.
[0025] In a preferred embodiment of a compressor according to the
invention, the regulation device comprises a throttling location
having an adjusting member. The throttling location can be,
especially, a throttling valve or a throttling flap. Also feasible
is a pressure reducer. The adjusting member provides for regulation
of the coolant mass flow or, that is, the inlet pressure acting on
the piston (P.sub.V1*). Such a measure is simple to put into
practice in terms of construction and ensures low manufacturing
costs.
[0026] Optionally, the adjusting member of the regulation device
adjusts the coolant mass flow or, that is, the inlet pressure in
dependence on the speed of rotation. Speeds of rotation are a
readily available regulation variable. Detection of speeds of
rotation can be accomplished, for example, by means of the
generation of electric pulses (induction principle); however,
direct, centrifugal-force-dependent regulation, for example, is
also feasible. Accordingly, regulation that is dependent on speed
of rotation opens up the possibility of many different forms of
construction for a compressor according to the invention, with low
manufacturing costs providing an advantage here too.
[0027] The throttling location preferably comprises an end-stop,
associated with the adjusting member, for a position of minimum
flow cross-section, this end-stop being so arranged that, even in
the case of very high speeds of rotation of the compressor, a
predetermined minimum coolant mass flow or, that is, inlet
pressure, is ensured. By this means, it is ensured in simple manner
that, even in the case of very high speeds of rotation, the
compressor does not automatically down-regulate completely as a
result of the throttling location or, that is, the additional
regulation device.
[0028] In a variant of a compressor according to the invention that
is simple and therefore economical to produce, the adjusting member
is an adjusting piston, which can especially be in the form a
stepped piston.
[0029] As an alternative to or in addition to the described
regulation device, the said control device is provided for control
or, that is to say, limitation of the inlet-gas-side coolant mass
flow or, that is, of the inlet pressure or, that is, the inlet
density. A control device usually has an arrangement that is simple
in terms of construction. In a preferred embodiment, the control
device comprises at least one inlet valve arranged on the inlet gas
side. The control device can be integrated in the inlet valve(s).
Preferably, the inlet valve is a pressure-controlled reed valve,
which in a variant that is simple in terms of construction is
formed by a valve plate, which has a throttling through-bore, and
an inlet blade. The inlet blade is preferably of tongue-like
construction.
[0030] When the compressor according to the invention is a
compressor having pistons, especially an axial piston compressor,
which has a cylinder block and at least one, but especially 5 to 9,
piston(s), which is/are axially movable back and forth in bores
provided in the cylinder block, an inlet valve can optionally be
associated with each cylinder, in which case the corresponding
inlet blades for the cylinders are integrated in an inlet blade
plate. This reduces the number of individual parts required for a
compressor according to the invention, which reduces the
manufacturing costs. That end of the or each cylinder space which
is associated with, or which faces, the inlet valve has, in a
further preferred embodiment, an annular extension which can be
bevelled off or flattened off towards the fixing location of the
inlet blade. By that means the stroke of the inlet blade can be
effectively limited.
[0031] Again assuming that the compressor is a compressor having
pistons, it is advantageous if the ratio of piston diameter and
piston stroke is approximately from 0.4 to 1.5, especially
approximately from 0.65 to 1.1. The ratio of piston diameter and
the throttling through-bore in the valve plate preferably is
approximately from 1.5 to 5, especially from 2.5 to 4. The ratio of
the throttling through-bore in the valve plate and the stroke of
the inlet blade is, in a further preferred embodiment,
approximately from 2.5 to 8, especially from 3.7 to 6.7. The ratio
of piston stroke to the stroke of the inlet blade can be
approximately from 10 to 30, especially from 14 to 24. The
above-described ratios, that is to say therefore the geometric
characteristics of the above-described compressor, are advantageous
in energy terms especially for compressors having CO.sub.2 as
coolant. Alternatively or additionally, the control device can also
comprise, arranged on the inlet gas side, an orifice plate which
appropriately defines the coolant mass flow or, that is, the inlet
pressure or, that is, the inlet density.
[0032] When the compressor according to the invention is a
compressor which has a tilt plate, the deflection angle of the tilt
plate, which angle governs the piston stroke of the compressor, is
governed to a very large extent by the interaction of, on the one
hand, the pressure P.sub.C in a drive mechanism chamber
substantially accommodating the tilt plate and, on the other hand,
the coolant mass flow on the inlet side or, that is, the inlet
pressure P.sub.V1*. A further force acting on the piston is
produced by the pressure P.sub.V2 on the high-pressure side. As a
result of regulation of the pressure in the drive mechanism
chamber, on the one hand, and regulation or control of the inlet
pressure P.sub.V1*, on the other hand, it is accordingly possible
for ideal regulation of the piston stroke to be accomplished, with
preference being given to modification of the pressure in the drive
mechanism compartment or, that is to say, the drive mechanism
chamber for "major" interventions for the purpose of regulation,
whereas a fine adjustment can be carried out by a rapid regulation
or, that is to say, the limitation, defined by the control device,
of the inlet pressure. As already mentioned, regulation or control
of the inlet pressure is associated with a substantially lower load
for the engine than regulation of the pressure in the drive
mechanism chamber, so that small rapid interventions for the
purpose of regulation can be carried out without major loss of
performance or do not even become necessary in the first place.
[0033] The regulation device can be actuatable or controllable from
outside the compressor. For the purpose there especially comes into
consideration a solenoid or like arrangement. This ensures simple
maintenance and simple replacement of the actuating device for the
regulation device.
[0034] In an embodiment which is simple in terms of construction
and which therefore is preferred, the regulation device and/or
control device include(s) an inlet-gas-side oil separator, which
has multifunctional significance. On the one hand, oil present in
the inlet gas is separated out; on the other hand, pressure
regulation or, that is, regulation of the coolant mass flow can be
simultaneously accomplished by that means.
[0035] In a variant of a compressor according to the invention, the
regulation device is self-regulating and, especially, dependent on
the difference in pressures at, on the one hand, the outlet side
or, that is to say, the high-pressure side and, on the other hand,
the entry side or, that is to say, the inlet side of the
compressor. This ensures reliable regulation of the compressor
taking into account the most important operating parameters.
[0036] Preference is given, especially in the case of a compressor
of the tilt plate form of construction, to the moment distribution
of those components of the compressor that are moved or movable in
rotation and in translation being such that, in the case of an
increase in the compressor speed of rotation, a substantially
constant regulation characteristic is ensured (that is say, the
moments are balanced). Explaining this again using the example of a
tilt plate compressor, this means that the tilt angle of the tilt
plate remains substantially the same or decreases. Accordingly, in
an advantageous combination of features of a compressor according
to the invention (again explained using the example of the tilt
plate compressor), there are three mechanisms which influence
regulation of the compressor, namely firstly the coolant mass flow
at the inlet gas side or, that is, the inlet pressure P.sub.V1*,
which is arranged to be controlled or regulated in accordance with
the invention, and also the pressure in the drive mechanism chamber
and the moment distribution of the components of the compressor.
This means that, by virtue of the form of construction of the tilt
plate and/or of the pistons, the moment distribution or, that is,
the ratio of moments causes the compressor to behave neutrally in
relation to the speed of rotation, that is to say especially it
does not up-regulate. Down-regulating action is, if required,
assisted by means of an appropriate regulating intervention for the
pressure P.sub.C in the drive mechanism chamber, with it being
possible especially for minor regulating interventions to be
accomplished without appreciable loss of engine performance by
means of adjustment or, that is to say, defined control/limitation
of the inlet pressure P.sub.V1*.
[0037] The regulation device and/or control device can be arranged
in an inlet gas channel extending mainly in the cylinder head of
the compressor. The inlet gas channel connects an inlet gas entry
of the compressor with an inlet gas chamber which is arranged
upstream of or, that is to say, before the inlet openings of the
individual cylinders.
[0038] The regulation device can furthermore comprise means for
measuring the coolant mass flow and/or the pressure in the inlet
gas channel (both to each side of the throttling location and also
to just one side of the throttling location) and/or the speed of
rotation of the compressor and/or of the engine driving the latter
and/or the pressure on the outlet side of the compressor. Depending
on the specific form of construction of the compressor, regulation
thereof is accordingly possible on the basis of a readily available
variable.
[0039] Optionally, the adjusting member of the regulation device
acts against the force of a resilient element, especially against
the force of a spring. This makes possible a wide range of
different regulation characteristics (depending on which
characteristic the resilient element has) and at the same time
constitutes automatic regulation of the adjusting member which is
simple to put into practice. In a preferred arrangement thereof,
the force of the resilient element can be adjusted especially by an
adjusting screw or like arrangement. This ensures that, with one
and the same construction, various regulation characteristics can
be set by simple means. Also, tolerances in the manufacture of the
resilient element and/or in the properties of the resilient element
can accordingly be compensated by simple means because fine
adjustment of the characteristic of the resilient element is
possible.
[0040] In a form of construction which is especially simple to put
into practice in constructional terms, the adjusting member is
arranged between the pressure gas side and inlet gas side and is
accordingly regulated automatically (where applicable against the
action of the resilient element) as a result of being subjected to
the pressures of, on the one hand, the pressure gas side and, on
the other hand, the inlet gas side.
[0041] In a further preferred embodiment, the regulation device
and/or control device comprise(s) a throttling location of constant
cross-section. This throttling location can be present, for
example, as the sole regulating device in a compressor according to
the invention or also in combination with a throttling location
which has an adjusting member. Especially in the case of
compressors with a high working pressure, that is to say, for
example, compressors which use CO.sub.2 as coolant, it is possible
just by means of this simple constructional measure for the desired
outcome to be achieved. Optionally, the inlet line and/or a
connection between the compressor and an evaporator are a component
part of the regulation device, it being possible, especially in the
case of a throttling location of constant cross-section, for
efficient regulation of the compressor to be achieved by means of
appropriate design of the inlet line and/or of the connection
between the compressor and the evaporator.
[0042] An especially efficient variant of a compressor according to
the invention is produced when across the regulation device there
is a pressure difference of approximately 1 bar at a compressor
speed of rotation of approximately 600 rpm and/or of approximately
10 bar at approximately 8000 rpm. The regulation device or, that
is, the throttling location preferably comprises a tubular line
having a tubular cross-section of approximately from 8 mm to 10 mm,
which ensures, especially in the case of a throttling location of
constant cross-section, a desired regulation characteristic. As
already mentioned hereinbefore, CO.sub.2 is used as coolant in a
particular form of construction of a compressor according to the
invention.
[0043] A compressor which is especially efficient and which manages
with few interventions for the purpose of regulation relating to
the pressure P.sub.C in the drive mechanism chamber is produced
when the moment due to those components of the compressor that are
moved in rotation is of substantially equal magnitude to the moment
due to those components of the compressor that are moved in
translation, that is to say when the behaviour of the compressor is
neutral, in terms of its regulation behaviour, with regard to speed
of rotation.
[0044] The invention will be described hereinbelow with reference
to further advantages and features by way of example and with
reference to the accompanying drawings. The drawings show in
[0045] FIG. 1 a first preferred embodiment of a compressor
according to the invention in a sectional view,
[0046] FIG. 2 a diagram of the mode of operation of a second
preferred embodiment of a compressor according to the
invention,
[0047] FIG. 3 a diagram of a co-ordinate system on which the
calculation of the moment ratios is based,
[0048] FIG. 4 a sectional view and an exploded view of a tilt plate
mechanism;
[0049] FIGS. 5 to 7 the regulation characteristics of a compressor
for various moment ratios of the components of the compressor that
are movable in rotation and that are movable in translation,
[0050] FIG. 8 a piston of a compressor of the first or second
preferred embodiment with the pressure conditions acting on it,
[0051] FIGS. 9a to 9c a mass flow diagram, a p-V diagram and a
regulation characteristic of a compressor according to the
invention,
[0052] FIGS. 10 to 13 further examples of regulation
characteristics of a compressor which has an up-regulating tendency
in the case of increasing speed of rotation and of a compressor
which exhibits a down-regulating tendency, and
[0053] FIG. 14 a detail, in diagrammatic form, of a third preferred
embodiment of a compressor according to the invention.
[0054] From FIG. 1 it can be seen that a first preferred embodiment
of a compressor according to the invention comprises a housing 1, a
cylinder block 2 and a cylinder head 3. Pistons 4 are mounted in
the cylinder block 2 so as to be movable axially back and forth.
The compressor is driven by means of a drive shaft 6, from a belt
disc 5. The compressor in this case is a compressor having variable
piston stroke, the piston stroke being governed by the deflection
angle of a tilt plate 7. The tilt plate 7 is in operative
engagement with the pistons 4 by way of sliding blocks 8 and is
driven in rotation by the drive shaft 6. The deflection angle of
the tilt plate 7 can, as is known from the prior art, be influenced
by a pressure change in a drive mechanism chamber 8', in which the
tilt plate is substantially arranged. The drive mechanism chamber
8' can be subjected to pressures between an inlet pressure, that is
to say the pressure prevailing on the inlet side of the compressor,
and a high pressure, that is to say the pressure prevailing on the
outlet side of the compressor. Depending on the pressure prevailing
in the drive mechanism chamber 8' or, that is, depending on the
difference in the pressures on the inlet side and in the drive
mechanism chamber, there is produced a predetermined deflection
angle for the tilt plate and, accordingly, a predetermined pressure
on the outlet side of the compressor.
[0055] A second quantity influencing the deflection angle of the
tilt plate 7 is the distribution of moments between the components
of the cylinder that are movable in translation, for example the
piston 4 or the sliding blocks 8, and the components of the
compressor that are movable in rotation, for example the tilt plate
7. In this case, by means of an appropriate mass or, that is,
moment distribution it is possible to achieve a rather
down-regulating tendency for the compressor at high speeds of
rotation. This is desirable, particularly in the case of modern
compressors, in order to be able to avoid icing-up, especially at
high speeds of rotation, without a large number of interventions
for the purpose of regulation. More precise details of the exact
constructional arrangement relating to the moments will be given
following a brief explanation of the further important features of
the compressor according to the invention in accordance with the
first preferred embodiment.
[0056] As can furthermore be seen in FIG. 1, an inlet gas channel 9
is arranged in the cylinder head 3, which channel connects an inlet
gas entry 10 to an inlet gas chamber 11, which is arranged upstream
of the cylinders. The compressed fluid or, that is to say, coolant
is made available to the coolant circuit by way of a pressure gas
chamber or, that is to say, outlet chamber 12. For regulation of
the coolant mass flow on the inlet gas side and also therefore of
the pressure on the inlet side of the compressor, a regulation
device is arranged in the inlet gas channel 9. This regulation
device comprises an adjusting piston 13 (which as an alternative to
the arrangement shown can also be in the form of a stepped piston),
a resilient element in the form of a spring 14 and also an
adjusting screw 15, which serves to adjust the biasing of the
spring 14. On its side facing the outlet chamber or, that is to
say, the pressure gas chamber 12, the adjusting piston 13 is
subject to the outlet pressure or, that is to say, the high
pressure, whereas on its side facing the adjusting screw 15, that
is to say on the side facing the inlet gas entry 10, it is
subjected to the inlet pressure or, that is to say, the entry
pressure. The higher the outlet pressure of the compressor, the
further the piston 13, which constitutes the adjusting member of
the regulation unit, is pressed into the inlet gas channel 9 and,
as a result, narrows the cross-section thereof. This results in a
lower coolant mass flow into the inlet gas chamber 11, which leads
to a lower pressure in the inlet gas chamber 11 and accordingly in
down-regulating behaviour of the compressor.
[0057] It is not clearly visible from the drawing that an end-stop
for a position of minimum flow cross-section is associated with the
adjusting piston 13, thereby ensuring that, even in the case of
very high speeds of rotation of the compressor and a relatively
high output pressure, a predetermined minimum coolant flow or, that
is, inlet pressure is ensured. The regulation unit comprising the
adjusting piston 13, the spring 14 and the adjusting screw 15 is
accordingly self-regulating, the regulation being accomplished in
dependence on the pressures on the outlet side and on the entry
side or, that is to say, the inlet gas side. Accordingly it can at
this point be stated that the deflection angle of the tilt plate 7
is governed by the interaction of the pressure in the drive
mechanism chamber 8', on the one hand, and the coolant mass flow on
the inlet gas side or, that is, the inlet pressure, on the other
hand, the inlet pressure itself being dependent in turn on the
output pressure of the compressor, so that feedback regeneration is
brought about for the compressor.
[0058] As an alternative to the automatic regulation shown in FIG.
1, the regulation device which is indicated in the diagram of FIG.
2 in general terms as the throttling location 17 can of course also
be regulated by external regulating variables and also by external
apparatus, for example a solenoid. FIG. 2 shows that the throttling
location 17 or, that is to say, the throttle (adjusting member) is
regulated by an external signal 16. This signal can be generated,
for example, on the basis of a measurement of the mass flow, of the
pressure on the high-pressure side or of a pressure difference
between the inlet gas channel and the high-pressure side or, that
is to say, of a pressure difference in the inlet gas channel
obtained from the different pressures P.sub.V1 and P.sub.V1* across
the throttling location 17. Of course, other parameters, for
example a speed of rotation or also temperatures or like
quantities, are also feasible as the basis for the signal 16.
[0059] FIG. 2 also shows a diagrammatic representation of the
coolant circuit in an h vs. log p diagram (supercritical process,
with CO.sub.2 as coolant) in a representation at the throttling
location (.delta.PV).
[0060] As already mentioned in the description of FIG. 1, the
distribution of moments between the masses of the compressor that
are moved in translation, for example the pistons 4, and the masses
moved in rotation, which include, for example, the tilt plate 7,
also has a regulating effect on the compressor. This moment ratio
will be discussed in somewhat greater detail hereinbelow. For
illustrative purposes, a simplified derivation to be considered as
being given by way of example will be considered for the various
moments. In the case of complex geometries, especially of the tilt
plate (when the illustrative approach no longer provides
satisfactory results), the mass moments of inertia and moments of
deviation and also other variables influenced by the geometry and
density of the materials can be calculated in simple manner by
CAD.
[0061] In the simplified, yet illustrative derivation of the mass
moments of inertia it is assumed that the centre of gravity of the
tilt plate is located at the tilting articulation on the mid-axis
of the shaft, that is to say no Steiner component or the like has
to be taken into account. For the derivation of the moment of
deviation the following mathematical relationships generally apply
(the co-ordinate system on which it is based being shown in FIG.
3):
J.sub.yz2=-J.sub.3 cos .alpha..sub.2 cos .alpha..sub.3-J.sub.2 cos
.beta..sub.2 cos .beta..sub.3-J.sub.3 cos .gamma..sub.2 cos
.gamma..sub.3
.alpha..sub.1=0 .beta..sub.1=90.degree. Direction angles of the x
axis .gamma..sub.1=90.degree. relative to the main inertia axes
.xi., .eta., .zeta. .alpha..sub.2=90.degree. .beta..sub.2=.psi.
Direction angles of the y axis .gamma..sub.2=90.degree.+.psi.
relative to the main inertia axes .xi., .eta., .zeta.
.alpha..sub.390 .degree. .beta..sub.3=90.degree.-.psi. Direction
angles of the z axis .gamma..sub.3=.psi. relative to the main
inertia axes .xi., .eta., .zeta.
[0062] As mentioned hereinabove, the co-ordinate system used herein
can be seen from FIG. 3. The following also applies to a
"ring":
J 2 = J .eta. = m 4 ( r a 2 + r i 2 + h 2 3 ) and ##EQU00001## J 3
= J .zeta. = m 2 ( r a 2 + r i 2 ) ##EQU00001.2##
(Note: J.sub.3.apprxeq.2J.sub.2)
[0063] For the moment of deviation, which governs the tilting
movement, the following applies:
J.sub.yz=-J.sub.2 cos .psi. sin .psi.+J.sub.3 cos .psi. sin
.psi.
[0064] Independently of FIG. 3, the following holds true for the
moment due to mass forces of the pistons:
.beta. i = .theta. + 2 .pi. ( i - 1 ) 1 n ##EQU00002## Z i = R
.omega. 2 tan .alpha. cos .beta. i ##EQU00002.2## F m i = m k z i
##EQU00002.3## M ( F m i ) = m k R cos .beta. i z i ##EQU00002.4##
M k , ges = m k R i = 1 n z i cos .beta. i ##EQU00002.5##
and also the moment M.sub.SW due to the moment of deviation:
M sw = J yz .omega. 2 ##EQU00003## J yz = { msw 2 ( r a 2 + r i 2 )
- msw 4 ( r a 2 + r i 2 + h 2 3 ) } cos .alpha. sin .alpha.
##EQU00003.2## J yx = msw 4 sin 2 .alpha. ( 3 r a 2 + 3 r i 2 - h 2
) ##EQU00003.3##
[0065] In the context of the invention, the following moment ratio
should be established by structural means for any desired tilt
angle or tilt angle range:
M.sub.SW.gtoreq.M.sub.k,ges or, preferably, the sub-case
M.sub.SW=M.sub.k,ges
[0066] As a result, the following also applies:
[ .omega. z R 2 m k tan .alpha. i = 1 n cos 2 .beta. .ltoreq.
.omega. 2 msw 24 sin 2 .alpha. ( 3 r a 2 + 3 r i 2 - h 2 ) ]
##EQU00004##
[0067] As already mentioned, the (tilting) moment of the tilt plate
due to the associated moment of deviation can be deliberately
adjusted by means of various parameters (geometry, density
distribution, mass, mass centre of gravity) so that
M.sub.SW.gtoreq.M.sub.k,ges or, however, the sub-case
M.sub.SW=M.sub.k,ges.
[0068] In the context of the equations given, the variables denote
the following: [0069] .theta. rotation angle of the shaft (the
considerations above and below being made on the basis of .theta.=0
for the sake of simplicity) [0070] .eta. number of pistons [0071] R
distance from piston axis to shaft axis [0072] .omega. speed of
rotation of shaft [0073] .alpha. tilt angle of tilt ring/tilt plate
[0074] m.sub.k mass of a piston including sliding blocks or, that
is to say, pair of sliding blocks [0075] m.sub.k,ges mass of all
pistons including sliding blocks [0076] M.sub.SW mass of tilt ring
[0077] r.sub.a external radius of tilt ring [0078] r.sub.i internal
radius of tilt ring [0079] h height of tilt ring [0080] g density
of tilt ring [0081] V volume of tilt ring [0082] .beta..sub.i angle
position of piston i [0083] z.sub.i acceleration of piston i [0084]
F.sub.mi mass force of piston i (including sliding blocks) [0085]
M(F.sub.mi) moment due to mass force of piston i [0086] M.sub.k,ges
moment due to mass force of all pistons [0087] M.sub.SW moment due
to advancing moment of tilt ring/tilt plate as a result of the
moment of deviation (J.sub.yz)
[0088] FIG. 4 shows the drive mechanism of the tilt plate form of
construction used as the basis, by way of example, for the
derivation. In the derivation, the tilt moment M.sub.SW due to the
moment of deviation J.sub.yz of the tilt plate is, in simplified
manner, set against the masses moved in translation or, that is,
the moment M.sub.K,ges produced thereby. Forces and moments of the
pins and/or of the gas force support or the like are, in
simplifying manner, not included in the calculation scheme, being
of subordinate importance.
[0089] It can be seen from the mathematical relationships that the
influence of the speed of rotation can be reduced out from the
equation. Also otherwise included are geometric variables which are
in particular relationships to one another and which, including
component densities and density distributions, can in principle be
so selected that the sum of the moments due to mass forces can be
adjusted to zero.
[0090] FIGS. 5, 6 and 7 each include a calculation scheme in
accordance with the equations used. Also shown, as the calculation
result, is the moment equilibrium. In addition, there is shown a
(qualitative) tilt characteristic, as would result if the gas
forces were taken into account.
[0091] The tilt characteristics of FIGS. 5, 6 and 7 result when, in
addition to the variation of speed of rotation and drive mechanism
chamber pressure in addition to the described forces and moments, a
particular inlet pressure and a particular high pressure are
imposed for system-related reasons. In the process it is assumed
that the inlet pressure prevailing upstream of the compressor and
the high pressure prevailing downstream of the compressor
approximately correspond to the inlet pressure and the high
pressure in the compressor, that is to say no throttling takes
place in the compressor. With regard to the moment equilibriums
calculated in accordance with the given equations there is obtained
according to
[0092] FIG. 5 a drive mechanism having up-regulating behaviour
[0093] FIG. 6 a drive mechanism having down-regulating behaviour,
and
[0094] FIG. 7 a drive mechanism having neutral behaviour.
[0095] Using FIG. 7 and also the equation for the sum of moments,
the influence of the tilt angle can be analysed in simple manner.
The effect results from the plots of the terms tan(.alpha.) and
sin(2.alpha.). This means that in the calculation the sum of
moments can be balanced for precisely one tilt angle; if this is
done, for example, for the maximum tilt angle of the tilt plate,
there are relatively small deviations in the sum of moments for
other tilt angles. It is, however, possible to keep these
deviations relatively small.
[0096] It is feasible to set the moment equilibrium for the
following tilt angles:
for .alpha..sub.min<=.alpha.<=.alpha..sub.max:
M.sub.K,ges=M.sub.SW for
.alpha.=(.alpha..sub.min-.alpha..sub.max)/2: M.sub.K,ges=M.sub.SW
for .alpha..sub.max=.alpha..sub.max: M.sub.K,ges=M.sub.SW for
.alpha.>=.alpha..sub.max: M.sub.K,ges=M.sub.SW
[0097] The two last-mentioned cases are to be preferred.
[0098] The advantage of a drive mechanism which is substantially
balanced in terms of its sum of moments lies in the fact, inter
alia, that when the speed of rotation increases the piston stroke
does not increase as well, which is to say that in such a
(undesirable) case there would be two effects present which would
have to be compensated. It can accordingly be stated that the case
is to be preferred where M.sub.k,ges is approximately equal to
M.sub.SW, which results in the compressor's having regulation
behaviour which is neutral with respect to speed of rotation. If
desired, M.sub.SW can also be selected so as to be greater than
M.sub.k,ges, which results in down-regulating behaviour of the
compressor at high speeds of rotation; on no account, though, is
the case desirable where M.sub.k,ges is greater than M.sub.SW
(up-regulation of the compressor in the case of increasing speed of
rotation).
[0099] As mentioned hereinbefore, it is preferred for M.sub.k,ges
to correspond approximately to M.sub.SW. As can be seen from FIG.
7, in a plot of the drive mechanism chamber pressure against the
tilt angle, the course of the curves is very similar for all speeds
of rotation n, with approximate equality of moments. This is also
reflected in a plot of the moment against the tilt angle, from
which it can be seen that for all tilt angles the sum of the
moments is almost constant. The individual moments do certainly
vary, however, for different tilt angles, with M.sub.k,ges
increasing for greater tilt angles in the entire range shown
whereas M.sub.SW decreases for greater tilt angles, resulting in
the sum of moments M.sub.k,ges+M.sub.SW shown, which is
approximately constant. Accordingly, a compressor characterised by
a moment plot of such a kind is, in terms of its regulation
characteristic, almost independent of the speed of rotation.
[0100] If the effect due to the sum of moments has a
down-regulating action, at least the tendency is the right one.
However, the influence of the speed of rotation on the effective
moments M.sub.SW and M.sub.K,ges is quadratic, compared to the
linear influence of the speed of rotation on the stroke volume and
is accordingly only of qualified suitability for keeping the
delivered mass flow constant.
[0101] Given a drive mechanism that behaves neutrally in the case
of changes in the speed of rotation, it is substantially only as a
result of change in the pressures P.sub.V1 (inlet pressure),
P.sub.V2 (high pressure or, that is to say, outlet pressure) and
the drive mechanism chamber pressure P.sub.C that the tilt angle of
the tilt plate changes. At a constant operating point where
P.sub.V1 and P.sub.V2 are as prespecified, change occurs
substantially only as a result of the drive mechanism chamber
pressure P.sub.C.
[0102] When a drive mechanism is designed in accordance with the
described criteria, the behaviour in relation to the delivered
coolant mass flow is proportional when there is a change in the
speed of rotation. This means that if the compressor speed of
rotation doubles, with the tilt plate tilt angle remaining the
same, which is the case with a drive mechanism having neutral
behaviour, then approximately double the amount of coolant is
delivered. Delivery of precisely double the amount of coolant is
the result if further losses that occur as a result of changed flow
conditions etc. are disregarded. If the changed flow conditions are
taken into account, discrepancies may occur.
[0103] In order to keep the delivered coolant mass flow constant,
or to limit it, in the case of a, for example, substantial increase
in the speed of rotation, there is provided, as already described
in the description of FIGS. 1 and 2, in the region of the coolant
inlet, a throttling location which is variable and which provides
rapid intervention.
[0104] It is possible for the compressor to be so designed that the
throttling intervenes in direct dependence on the compressor speed
of rotation (as, for example, in the second preferred embodiment;
see FIG. 2). In the first preferred embodiment (see FIG. 1), on the
other hand, the cross-section of the throttling location is a
function of the high pressure P.sub.V2 of the compressor, that is
to say the throttling is controlled in dependence on the high
pressure.
[0105] When the compressor speed of rotation increases (for
example, suddenly), then the pressure P.sub.V2 increases
approximately just as quickly. Because no substantial throttling
occurs on the high-pressure side, P.sub.V2 can be assumed to be
both the system-side high pressure and also the pressure level on
the high-pressure side in the cylinder head. Pressure losses in the
pipework play just a subordinate part so that they can be
disregarded in this analysis. In the case of the mentioned increase
in the compressor speed of rotation the inlet pressure moreover
decreases, with the pressure P.sub.V1, which prevails upstream of
the throttling location, approximately maintaining its level
(system-side pressure level on the inlet side), whereas the
pressure P.sub.V1* downstream of the throttling location drops
compared to P.sub.V1. The pressure P.sub.V2 now acts, as a
fundamental adjustment variable (in addition to the inlet
pressure), on the throttling mechanism in such a way that the
cross-section of the throttling location is reduced.
[0106] Lowering of the pressure P.sub.V1 to the pressure P.sub.V1*
downstream of the throttling location has the consequence that a
reduced inlet density (reduced pressure) is applied to the
cylinders (at the inlet valves); as a result, the pressure in the
cylinder or, that is, at the end faces of the pistons (which are
directed towards the valve plate) decreases so that the tilt angle
has a tendency to reduce. This moreover results in the fact that
the pressure P.sub.V2 reduces again, which in turn results in
feedback to the throttle.
[0107] Because the inlet condition of the compressor, which can be
described by the variables t.sub.V1 and P.sub.V1 (see FIG. 14), is
substantially unchanged, the mentioned expansion valve will not
change its setting and the pressure levels P.sub.V1 and P.sub.V2
also remain the same. It should be mentioned at this point that the
regulation path can also be arranged differently so that, instead
of P.sub.V1, the pressure P.sub.V2 can be used for defining the
state of the compressor.
[0108] To summarise, the thermodynamic variables before the
compressor and after the compressor (in the direction of
circulation) remain the same and the regulation element does not
intervene in the system.
[0109] In addition to a thermostatic expansion element, differently
operating and differently actuated expansion elements are of course
also feasible.
[0110] As a result of the increase in the compressor speed of
rotation, the compressor regulates itself automatically by means of
the fact that, in addition to the drive mechanism chamber pressure
P.sub.C, the inlet pressure P.sub.V1 or the high pressure P.sub.V2
come to have a regulating effect. Because P.sub.V1 and P.sub.V2 are
also affected by the operating state of the system, where usually
it is not necessarily desirable also to have changes in the case of
a changed compressor speed of rotation, a pressure P.sub.V1* is
brought about which, being the gas force applied to the pistons,
can intervene in the force equilibrium or, that is, moment
equilibrium of the tilt plate.
[0111] This means that, in an operating state with a suddenly
increasing speed of rotation, the inlet pressure after the
throttling location is reduced so that the pressure level P.sub.V1*
is established, in order to keep the high pressure P.sub.V2 and the
mass flow at the same level. Acting on the pistons are, on one
side, the pressures P.sub.V1* and P.sub.V2 and also, on the other
side (on the drive mechanism side), the pressure P.sub.C (see FIG.
8). When P.sub.V1 is lowered to the level P.sub.V1*, the tilt angle
of the tilt plate is reduced, that being the case without having to
change the crank chamber pressure. This means that, in contrast to
the prior art, where the drive mechanism chamber pressure P.sub.C
is used as an adjusting variable, a further adjusting variable
P.sub.V1* is introduced in accordance with the invention.
[0112] The pressure P.sub.V1* can be substantially smaller than
P.sub.V1 (certainly by 5 to 15 bar). Because such throttling can,
depending on the operating point, be associated with substantial
losses, the throttling location or, that is to say, the regulation
device is variable over a wide range.
[0113] In a preferred embodiment, the throttling location, which
depending on its position constricts the inlet gas line
cross-section to a greater or lesser extent, has three different
operating ranges: [0114] In the first position, no throttling is
brought about (operating position 1). [0115] In the second
position, acting on the inlet side on the piston are a pressure
between the pressures P.sub.V1 and P.sub.V1* and on the
high-pressure side the pressure P.sub.V2. Also acting as a guiding
variable is, for example, the force of a pressure spring. In the
second position, substantial or less substantial, depending on the
gas forces applied, throttling occurs (operating position 2).
[0116] In a third position, the adjusting piston can, when the flow
cross-section in the inlet line has reached a minimum, hit an
end-stop. In that event, a minimum flow cross-section is maintained
(operating position 3).
[0117] It is to be understood that the principle proposed here is
to be regarded as being only by way of example. There can also be
used, for example, a stepped piston, where for the pressures
P.sub.V1 and P.sub.V2 there is available in each case a different
application diameter. It should be mentioned at this point that the
adjusting member or, that is to say, the piston should operate in
leakage-free manner as far as possible, which is ensured by means
of piston rings. Other sealing measures are also feasible.
[0118] FIG. 9a shows, for a prespecified pressure level P.sub.V1
and P.sub.V2 of the air-conditioning system, the adjustable mass
flows (qualitative representation), whilst FIG. 9b shows the p-V
diagram corresponding thereto. Starting from the origin, the
achievable mass flow increases along with the speed of rotation.
The envelope curve for the corresponding slope shows the delivered
mass flow for a maximum tilt plate tilt angle/maximum geometric
stroke volume. The coolant mass flow m.sub.1 at the speed of
rotation n.sub.1 doubles, for example, to a coolant mass flow
m.sub.2 in the event of a corresponding increase in the speed of
rotation of n.sub.2=2.times.n.sub.1. The greater the desired
coolant mass flow, the greater the compressor speed of rotation
also has to be in order to achieve that flow at maximum compressor
stroke. Once the desired coolant mass flow has been achieved, for
example m.sub.1, m.sub.2 or m.sub.3, no increase in the coolant
mass flow is desired in the event of further increase in the speed
of rotation. The horizontal plots for m.sub.1, m.sub.2 and m.sub.3
shown in the diagram correspond in each case to a particular drive
mechanism chamber pressure, which is approximately constant. In the
region of the horizontal lines, the effect of the inlet-gas-side
throttling location comes into play with increasing speed of
rotation. Whereas on the envelope curve the inlet-gas-side throttle
is in operating position 1 (no throttling), in operating range 2
the throttling cross-section is reduced with increasing speed of
rotation.
[0119] Consequently, when the throttling location is appropriately
designed for various coolant mass flows, which are established by
virtue of a particular operating state, the mass flow can be kept
constant.
[0120] When, for example, the operating state is established, by
means of P.sub.V1, t.sub.V1 and P.sub.V2, for the pressures on the
high-pressure side and on the inlet side of the system and also for
the inlet state at the compressor entry and, in that state of
operation, the speed of rotation n.sub.2 is present with a coolant
mass flow m.sub.1, the throttling location is in the operating
state 2, that is to say the inlet cross-section of the inlet line
is reduced in the region of the throttling location with respect to
the initial state (operating state 1). Furthermore, in addition to
the pressure level at the compressor entry P.sub.V1, an inlet
pressure P.sub.V1* will have been established, which because of the
throttling is lower than the pressure P.sub.V1. On further
reduction of the pressure P.sub.V1*, the gas forces acting on the
piston become lower so that, with the drive mechanism chamber
pressure remaining approximately the same, the tilt angle of the
tilt plate is reduced (in contrast to the prior art; see FIG. 9c in
this regard). A reduction in the tilt plate tilt angle leads in
turn to a lower mass flow. In this case, it is substantially not by
means of the fact that a pressure loss is produced or that the
volumetric efficiency or, that is to say, degree of filling becomes
poorer that the delivered amount is limited or kept constant;
mainly the pressure reduction intervenes directly in the
equilibrium of stroke adjustment and down-regulates the stroke in
the case of increasing speed of rotation. It should be mentioned at
this point that P.sub.V1* should not drop too substantially,
because otherwise excessive losses are brought about.
[0121] The regulation behaviour is especially characterised in
that, in contrast to the prior art, where given a constant
operating state of the system described by P.sub.V1, t.sub.V1 and
P.sub.V2 exactly one tilt plate tilt angle is associated with each
drive mechanism chamber pressure P.sub.C (see FIG. 5, although
exceptions occur in the case of very high speeds of rotation or
very small tilt angles (maxima), a plurality of tilt plate tilt
angles are feasible for a drive mechanism chamber pressure P.sub.C.
In contrast to the prior art, not only is P.sub.C an adjusting
quantity but so is the pressure P.sub.V1* too.
[0122] The pressure difference P.sub.V1*--P.sub.C can reach
negative values. In the case of the prior art P.sub.C-P.sub.V1 must
be used as the basis. The pressure P.sub.C is in that case always
higher than the pressure P.sub.V1. As a result, the regulation
range is also greater (.DELTA.p) in accordance with the
invention.
[0123] In conclusion it must be mentioned again that in addition to
the adjusting variable acting on the adjusting piston (the coolant
mass flow m or, that is, P.sub.V2), external actuation of an
adjusting piston or throttling device can also be carried out (by
means of a solenoid or the like; see FIG. 2). Such an arrangement
must be "informed" of the mass flow increase in the form of a
signal, for example by detecting the inlet-side or
high-pressure-side pressure difference (throttling
location/measurement orifice (variable or non-variable) on the
inlet side or high-pressure side of the compressor).
[0124] FIGS. 10 to 13 show continuations of the qualitative
representations of FIG. 6 and FIG. 7, with a drive mechanism that
is independent of the speed of rotation being shown in FIGS. 10 and
11 and with a drive mechanism that, analogously to FIG. 7, favours
down-regulation in the case of increasing speed of rotation being
shown in FIGS. 12 and 13. It is shown that in the case of an
unchanged sum of moments (ratio of M.sub.SW and M.sub.K,ges from
FIGS. 6 and 7) throttling on the inlet side (formation of P.sub.V1*
as opposed to P.sub.V1) which is dependent on mass flow, pressure
or speed of rotation brings about further separation of the
characteristic curves.
[0125] Finally, FIG. 14 shows, in diagrammatic form, a third
preferred embodiment of a compressor according to the invention.
The third preferred embodiment is a compressor which does not have
a regulation device but rather a control device for the inlet
pressure. This results in the fact that the described compressor is
very simple in terms of construction and also, therefore,
economical to manufacture. It should be pointed out at this point
that the control device of the third preferred embodiment can be
put into practice in a compressor together with a regulation device
for the inlet pressure. Alternatively, however, a construction
which has only a control device is also feasible. The third
preferred embodiment comprises, as do the other preferred
embodiments as well, a plurality of pistons 4, which are mounted in
the cylinder block 2 so as to be movable back and forth.
[0126] The third preferred embodiment has, instead of a regulatable
throttling device, a valve plate 18 having an inlet blade 21
arranged below it, at the inlet side for the inlet gas into the
cylinder space. The inlet blade 21 is of tongue-like construction
and serves for control of the inlet gas entry. When the gas is
compressed in the cylinder, the inlet blade 21 closes a throttling
through-bore 19, whereas when the inlet gas is being drawn in
(brought about by the lower pressure prevailing in the cylinder)
the inlet blade 21 moves in a downward direction through a stroke t
(indicated by arrows 20) and allows the coolant or, that is, the
inlet gas being drawn in to enter the cylinder through the
throttling through-bore 19.
[0127] The throttling through-bore 19 has a diameter d. By virtue
of the geometry of the inlet valve, that is to say especially by
virtue of the diameter d of the throttling through-bore 19, and the
compressor geometry, desirable lowering of the pressure P.sub.V1*
is brought about over wide operating ranges of the compressor
according to the invention. This can be achieved, for example, (in
the case of a compressor having CO.sub.2 as coolant) using the
following parameters for the compressor geometry: The stroke t of
the inlet blade 21 is between 0.9 and 1.2 mm, whereas the valve
plate 18 has a bore (throttling through-bore) whose diameter d is
between 4.5 and 6 mm. The values for the piston diameter are
approximately from 15 to 18 mm and the piston stroke is
approximately from 17 to 22 mm. The maximum stroke volume per
cylinder is from 3 ccm to 6 ccm. This results in variables
describing the geometry of the compressor that are advantageous in
energy terms which are a ratio of piston diameter and piston stroke
of approximately from 0.65 to 1.1, a ratio of piston diameter and
the throttling through-bore in the valve plate of approximately
from 2.5 to 4, a ratio of the throttling through-bore in the valve
plate and stroke of the inlet blade of approximately from 3.7 to
6.7 and a ratio of piston stroke to the stroke of the inlet blade
of approximately from 14 to 24. It should be noted at this point
that these values reflect the optimum geometry for operation with
CO.sub.2 as coolant but that, depending on constructional
requirements, values of from 0.4 to 1.5 for the ratio of piston
diameter and piston stroke and values of from 1.5 to 5 for the
ratio of piston diameter and the throttling through-bore and values
of from 2.5 to 8 for the ratio of the throttling through-bore and
stroke of the inlet blade and values of approximately from 10 to 30
for the ratio of piston stroke to the stroke of the inlet blade are
also advantageous in energy terms. Accordingly, in this preferred
embodiment, the throttling through-bore 19 on the inlet side is
used as a throttling location or, that is to say, control device
and is appropriately designed in conjunction with the other
parameters regulating the compressor. A construction of such a kind
is, especially, very effective in compressors which are already
optimised in terms of moments, that is to say which have an optimum
relationship between the moments due to the rotary masses and due
to the masses moved in translation. The gas flowing in flows
through an inlet chamber, which is located in the cylinder head 2,
at a pressure P.sub.V1 and is then, by way of the inlet valve,
which has, for example, the configuration described above,
introduced into the cylinder bore, where by virtue of the inlet
valve configuration the pressure P.sub.V1* is established, which
ensures optimum regulation behaviour of the compressor.
[0128] In summary it should be stated that the throttling of the
inlet pressure or, that is, of the coolant flow produces a
down-regulating effect which primarily results not from a lowering
of the inlet density but rather from the direct utilisation of the
prevailing pressure difference at the throttle for the purpose of
stroke volume adjustment. Adjustment of the throttle results in
adjustment of the pressure difference prevailing at the throttle
and, therefore, in adjustment of the stroke volume. Furthermore, a
modification of the volume flow results in a change in the
prevailing pressure difference and, therefore, in subsequent
regulation of the stroke volume.
[0129] It can furthermore be stated that the advantages of the
invention lie, inter alia, in the fact that the tilt plate
compressor [0130] reacts less sensitively, or hardly reacts, to
variations in the speed of rotation caused by the belt drive (drive
mechanism) [0131] the losses due to interventions for the purpose
of regulation between the pressure levels inlet pressure, high
pressure and drive mechanism chamber pressure are reduced [0132]
the speed of regulation is improved [0133] the coolant mass flow
can be kept constant in a wide speed of rotation range, and [0134]
the coolant mass flow can be limited at high speeds of
rotation.
[0135] Although the invention is described using embodiments having
fixed combinations of features, it nevertheless also encompasses
any further feasible advantageous combinations of those features,
as are especially but not exhaustively mentioned in the subordinate
claims. All features disclosed in the application documents are
claimed as being important to the invention insofar as they are
novel on their own or in combination compared with the prior
art.
LIST OF REFERENCE NUMERALS
[0136] 1 housing [0137] 2 cylinder block [0138] 3 cylinder head
[0139] 4 piston [0140] 5 belt disc [0141] 6 drive shaft [0142] 7
tilt plate [0143] 8 sliding block [0144] 8' drive mechanism chamber
[0145] 9 inlet gas channel [0146] 10 inlet gas entry [0147] 11
inlet gas chamber [0148] 12 output gas chamber or pressure gas
chamber [0149] 13 adjusting piston [0150] 14 spring [0151] 15
adjusting screw [0152] 16 external signal [0153] 17 regulation
device [0154] 18 valve plate [0155] 19 bore [0156] 20 arrow [0157]
21 inlet blade [0158] 101 compressor [0159] 102 gas
cooler/liquefier [0160] 103 expansion element [0161] 104 evaporator
[0162] 105 regulation path
* * * * *