Actuation system for a resonant linear compressor, method for actuating a resonant linear compressor, and resonant linear compressor

Abstract

An actuation system for a resonant linear compressor (50), applied to cooling systems, the latter being particularly designed to operate at the electromechanical frequency of said compressor (50), so that the system will be capable of raising the maximum power supplied by the linear actuator, in conditions of overload of said cooling system. Additionally, an actuation method for a resonant linear compressor (50) is disclosed, the operation steps of which enable one to actuate the equipment at the electromechanical resonance frequency, as well as to control the actuation thereof in over load conditions.

Claims

1. An actuation system for a resonant linear compressor (50), the resonant linear compressor (50) being an integral part of a cooling circuit, the resonant linear compressor (50) comprising at least one cylinder (2), at least one head (3), at least one electric motor and at least one spring, the cylinder (2) housing a piston (1) operatively, the actuation system comprising at least one electronic actuation control (20) for actuating the electric motor, the electronic actuation control (20) comprising at least one control circuit (24) and at least one actuation circuit (26), associated to each other, the electronic actuation control (20) being electronically associated to the electric motor of the linear compressor (50), wherein the actuation system detects at least one overload condition of the linear compressor (50), through at least one electric magnitude measured or estimated by the electronic actuation control (20), and in response to the detected overload condition the actuation system implements an overload control mode in which an actuation frequency of the electric motor is set to match an electromechanical resonance frequency of the linear compressor until said overload condition subsides, wherein the electric magnitude measured or estimated is given by at least one of: (i) a piston velocity value (V.sub.p) or (ii) a piston displacement value.

2. The actuation system according to claim 1, wherein the overload control adjusts the actuation frequency of the electric motor based on the piston displacement value with respect to a maximum reference displacement (D.sub.REF).

3. The actuation system according to claim 1, wherein the overload control mode adjusts the actuation frequency of the electric motor based on a velocity phase value (φ.sub.v) of the motor of the compressor (50) with respect to a reference velocity phase (ϕ.sub.REF).

4. The actuation system according to claim 1, wherein the overload control mode adjusts the actuation frequency of the electric motor based on a displacement phase value (φ.sub.d) of the motor of the compressor (50) with respect to a reference displacement phase (φd.sub.REF).

5. The actuation system according to claim 1, wherein the overload control mode adjusts the actuation frequency of the electric motor based on a minimum current phase value (φ.sub.c).

6. A resonant linear compressor (50) comprising: a cylinder (2) operatively housing a piston; a head (3); an electric motor; a spring; and, an actuation system comprising an electronic actuation control (20) for actuating the electric motor, the electronic actuation control (20) comprising at least one control circuit (24) and at least one actuation circuit (26), associated to each other, the electronic actuation control (20) being electronically associated to the electric motor of the linear compressor (50), wherein the actuation system detects an overload condition of the linear compressor (50), through at least one of: (i) a piston velocity value (V.sub.p) of the piston as determined by the electronic actuation control (20); or (ii) a piston displacement value (d.sub.p) of the piston as determined by the electronic actuation control (20); and wherein said actuation system comprises an overload control mode that is initiated when said actuation system detects said overload condition and in which an actuation frequency of the electric motor is adjusted and maintained by said electronic actuation control during said overload condition to be the same as an electromechanical resonance frequency of said linear compressor until said overload condition subsides such that a maximum power supplied by said linear compressor is increased during said overload condition.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

(1) The present invention will now be described in greater details with reference to the attached drawings, in which:

(2) FIG. 1 represents a schematic view of a resonant linear compressor;

(3) FIG. 2 illustrates a schematic view of the mechanical model of the resonant linear compressor employed in the present invention;

(4) FIG. 3 illustrates a schematic view of the electric model of the resonant linear compressor of the present invention;

(5) FIG. 4 shows a graph of the position of the poles of the electric, mechanical and complete system, according to the teachings of the present invention;

(6) FIG. 5 illustrates a Bode diagram for the displacement of the mechanical system;

(7) FIG. 6 shows a Bode diagram for the velocity of the mechanical system;

(8) FIG. 7 illustrates a Bode diagram of the current of the complete electromechanical system of the present invention;

(9) FIG. 8 illustrates a Bode diagram of the displacement of the complete electromechanical system, according to the teachings of the invention;

(10) FIG. 9 illustrates a Bode diagram of the velocity of the complete electromechanical system of the present invention;

(11) FIG. 10 represents a simplified block diagram of the control with a sensor;

(12) FIG. 11 illustrates a block diagram of the control and of the inverter with a sensor;

(13) FIG. 12 shows a simplified block diagram of the control without sensor;

(14) FIG. 13 shows a block diagram of the control and inverter without sensor;

(15) FIG. 14 shows first flow chart capable of detecting the overload mode in a normal control proposal;

(16) FIG. 15 shows second flow chart intended for detection of the overload mode in a second normal control proposal;

(17) FIG. 16 shows an overload-control flow chart for maximum displacement;

(18) FIG. 17 shows an overload-control flow chart for the adjustment of the velocity phase;

(19) FIG. 18 shows an overload-control flow chart for the adjustment of the displacement phase; and

(20) FIG. 19 shows an overload-control flow chart for minimum current shift.

DETAILED DESCRIPTION OF THE FIGURES

(21) FIG. 1 shows a schematic view of a resonant linear compressor 50, object of the present invention.

(22) FIGS. 2 and 3 illustrate a mechanical and electric model of the linear compressor 50, such a mechanical model being defined on the basis of equation 1 below, and said electric model being defined from equation 2.

(23) $\begin{array}{cc}m.Math.\frac{d^{2}d\left(t\right)}{d{t}^2}=F_{\mathrm{MT}}\left(i\left(t\right)\right)-F_{\mathrm{ML}}\left(d\left(t\right)\right)-F_{\mathrm{AM}}\left(v\left(t\right)\right)-F_G\left(d\left(t\right)\right)& \left(1\right)\end{array}$
wherein: F.sub.MT(i(t))=K.sub.MT.Math.i(t)—motor force [N]; F.sub.ML(d(t))=K.sub.ML.Math.d(t)—spring force [N]; F.sub.AM(v(t))=K.sub.AM.Math.v(t)—damping force [N]; F.sub.G(d(t))—force of gas pressure in the cylinder [N]; K.sub.MT—motor constant K.sub.ML—spring constant K.sub.AM—damping constant m—mass of the moveable par v(t)—piston velocity d(t)—piston displacement i(t)—motor current
V.sub.ENT(t)=V.sub.R(i(t))+V.sub.L(i(t))+V.sub.MT(v(t))  (2)
Wherein: V.sub.R(i(t))=R.Math.i(t)—resistance voltage [V];

(24) $V_L\left(i\left(t\right)\right)=L.Math.\frac{di\left(t\right)}{dt}-\mathrm{inductor}\mathrm{voltage}\left[V\right];$ V.sub.MT(v(t))=K.sub.MT.Math.v(t)—voltage induced in the motor or CEMF [V]; V.sub.ENT(t)—feed voltage [V]; R—electric resistance of the motor L—motor inductance.

(25) It should be pointed out that, the gas pressure force (F.sub.G(d(t))) is variable with the suction and discharge pressures, with the non-linear piston displacement, with the other forces in the mechanical equation they are all linear, just as all the voltages in the electric equation. In order to obtain the complete model of the system, it is possible to replace the pressure force by the effects which it causes in the system, which are power consumption and variation in the resonance frequency.

(26) The power consumption may be modeled by an equivalent damping and the variation in the resonance frequency by an equivalent spring.

(27) Thus, the equation (1) above may be rewritten as follows:

(28) $\begin{array}{cc}m.Math.\frac{d^2\left(t\right)}{d{t}^2}=K_{\mathrm{MT}}.Math.i\left(t\right)-\left(K_{\mathrm{ML}}+K_{\mathrm{MLEq}}\right).Math.d\left(t\right)-\left(K_{\mathrm{AM}}+K_{\mathrm{AMEq}}\right).Math.v\left(t\right)\mathrm{or}& \left(3\right)\\ m.Math.\frac{d^2\left(t\right)}{d{t}^2}=K_{\mathrm{MT}}.Math.i\left(t\right)-K_{\mathrm{MLT}}.Math.d\left(t\right)-K_{\mathrm{AMT}}.Math.v\left(t\right)& \left(4\right)\end{array}$
Wherein: K.sub.MLEq—equivalent spring coefficient K.sub.AMEq—equivalent damping coefficient K.sub.MLT=K.sub.ML+K.sub.MLEq—total spring coefficient K.sub.AMT=K.sub.AM+K.sub.AMEq—total damping coefficient

(29) Applying the Laplace transform to the equations (2) and (4), one can obtain the equation (5) below, which represents the electric equation at the minimum of the frequency and the mechanical equations (6) and (7), which represent, respectively the function of transfer between displacement and velocity with the current.

(30) $\begin{array}{cc}I\left(s\right)=\frac{V_{\mathrm{ENT}}\left(s\right)-K_{\mathrm{MT}}.Math.V\left(s\right)}{L.Math.s+R}& \left(5\right)\\ \frac{D\left(s\right)}{I\left(s\right)}=\frac{K_{\mathrm{MT}}}{m.Math.s^2+K_{\mathrm{AMT}}.Math.s+K_{\mathrm{MLT}}}& \left(6\right)\\ \frac{V\left(s\right)}{I\left(s\right)}=\frac{K_{\mathrm{MT}}.Math.s}{m.Math.s^2+K_{\mathrm{AMT}}.Math.s+K_{\mathrm{MLT}}}& \left(7\right)\end{array}$

(31) The equation (8) below represents the characteristic equation of the electric system, so that the equation (9) represents the characteristic equation of the mechanical system. The poles of this equation define the mechanical resonance frequency, region where the relationship between displacement/current, or velocity/current, is maximum, and therefore with maximum efficiency as well, just as described ion other solutions of the prior art.
EC.sub.E=L.Math.s+R  (8)
EC.sub.M=m.Math.s.sup.2+K.sub.AMT.Math.s+K.sub.MLT  (9)

(32) Working out mathematically the equations (5) to (9), one can obtain the equations (10), (11) and (12), which represent, respectively, the function of transfer of the current, of the displacement and of the velocity of the piston of the compressor 50, as a function of the input voltage, for the complete electromechanical system, according to the teachings of the present invention:

(33) $\begin{array}{cc}\frac{I\left(s\right)}{V_{\mathrm{ENT}}\left(s\right)}=\frac{{\mathrm{EC}}_M}{{\mathrm{EC}}_M.Math.{\mathrm{EC}}_E+K_{\mathrm{MT}}^2.Math.s}& \left(10\right)\\ \frac{D\left(s\right)}{V_{\mathrm{ENT}}\left(s\right)}=\frac{K_{\mathrm{MT}}}{{\mathrm{EC}}_M.Math.{\mathrm{EC}}_E+K_{\mathrm{MT}}^2.Math.s}& \left(11\right)\\ \frac{V\left(s\right)}{V_{\mathrm{ENT}}\left(s\right)}=\frac{K_{\mathrm{MT}}.Math.s}{{\mathrm{EC}}_M.Math.{\mathrm{EC}}_E+K_{\mathrm{MT}}^2.Math.s}& \left(12\right)\end{array}$

(34) One may further define the equation (13) or (14) below, as the characteristic equation of the electromechanical system designed in the present invention:
EC.sub.S=EC.sub.M.Math.EC.sub.E+K.sub.MT.sup.2.Math.s  (13)
or:
EC.sub.S=m.Math.L.Math.s.sup.3+(K.sub.AMT.Math.L+m.Math.R).Math.s.sup.2+(K.sub.MLT.Math.L+K.sub.AMT.Math.R+K.sub.MT.sup.2).Math.s+K.sub.MLT.Math.R  (14)

(35) The pair of complex poles of the characteristic equation of the electromechanical system above defines the electromechanical resonance frequency, the region in which one has greater relation between current, the displacement and the velocity with the input voltage. Therefore, this is a region where it is possible to obtain maximum power of the resonant linear compressor, as proposed in the present invention.

(36) For a better understanding of the characteristics of the actuation system and method proposed, which will be described in greater details later, one presents the values in Table 1 below, which define the coefficients of a resonant linear compressor, designed to operate at a mechanical resonant frequency of 50 Hz, for a nominal load of 50 W.

(37) TABLE-US-00001 TABLE 1 Coefficients of the resonant linear compressor Coefficient Value Unit R 12.9 {acute over ( )}Ω L 0.75 H K.sub.MT 70 V .Math. s/m or N/A K.sub.MLT 81029.5 N/m K.sub.AMT 10 N .Math. s/m m 0.821 Kg

(38) Calculating the poles of the electric system and mechanical system in isolation, and of the complete electromechanical system, one will visualize the alteration in the system poles, according to Table 2 below, and also from FIG. 4.

(39) The mechanical resonance frequency is given by the module of the pair of complex poles of the characteristic equation of the mechanical system (314.2 rad/s or 50 Hz). The electromechanical resonance frequency is given by the module of the pair of complex poles of the characteristic equation of the electromagnetic system (326.6 rad/s or 51.97 Hz).

(40) TABLE-US-00002 TABLE 2 Poles of the electric, mechanical and electromechanical system Poles System Real Complex Electric  17.2 — Mechanical — 6.09 ± 3141j Electromechanical −15.9  6.73 ± 326.5j

(41) In Bode diagrams of the transfer function of displacement and velocity, for the mechanical system, such as shown in FIGS. 5 and 6, one can observe that, at the mechanical resonance frequency, the gain is maximum. In this case, the phase between the displacement with the current is of −90 degrees (displacement and current are in quadrature), and the phase of the velocity with the current is zero degree (velocity and current are in phase).

(42) Additionally, one observes from the diagrams of FIGS. 7, 8 and 9, represent, respectively, the Bode diagrams of the transfer functions of the current, the displacement of the velocity, as a function of the input voltage, which, at the electromechanical resonance frequency, the gain is maximum, according to the teachings of the present invention.

(43) Moreover, it is possible to observe, in FIG. 7, that, in the mechanical resonance frequency, the value of the current is minimum, for which reason the efficiency is maximum. At the middle point between the mechanical resonance frequency and the electromechanical resonance frequency, the power factor of the linear actuator is maximum, since the phase of the current has the shortest delay.

(44) The electromechanical resonance frequency is always above the mechanical resonance frequency, and at the electromechanical frequency the phase between the displacement and the input voltage is around −176 degrees, and the phase between the velocity and the input voltage is around −86 degrees, for the data presented in Table 1 above. The greater the difference between the real pole and the module of the pair of complex poles of the electromechanical system, the shift of the displacement and of the velocity will tend to −180 degrees and −90 degrees, respectively.

(45) In the face of the foregoing, one proposes the present invention for the main purpose of supplying maximum power to the resonant linear compressor 50, for conditions of overload of the cooling system.

(46) Such a system takes into account that the linear compressor 50 comprises at least one cylinder 2, at least one had 3, at least one electric motor and at least one spring, so that the cylinder 2 houses operatively a piston 1. FIG. 1 shows said compressor 50 and its constituent parts.

(47) As far as the electronic composition is concerned, it is possible to note, on the basis of FIGS. 10-13, the main characteristics of the present actuation system. Such a system comprises at least one electronic actuation control 20 of the electric motor, this electronic actuation control 20 being provided with at least one control circuit 24 and at least one actuation circuit 26, associated electrically with each other.

(48) The same figures show that the electronic actuation control 20 is electronically associated to the electric motor of the linear compressor 50, this electronic control 20 being composed of rectifying element, inverter (inverting bridge) and digital processor.

(49) A quite relevant characteristic of the presently claimed invention as compared with the prior techniques refers to the fact that the actuation system is particularly configured to detect at least one overload condition of the linear compressor (50), through at least one electric magnitude measured or estimated by the electronic actuation control 20, and to adjust, from an overload control mode, the actuation frequency of the electric motor to an electromechanical resonance frequency.

(50) The electric magnitude measured or estimated is given by a actuating piston velocity value V.sub.p, or still by a piston displacement value d.sub.p. The actuation electronic control 20 is capable of actuating, according to the teachings of the invention, the electric motor of the compressor 50 with a PWM senoidal voltage starting from an amplitude and a controlled range.

(51) As already mentioned before, the present invention has the central objective of detecting a condition of overload of the linear compressor 50, under conditions in which it is necessary to adjust the actuation frequency of said electric motor, in a determined operation mode in overload, in order to achieve the desired control of the cooling system in situations of high demand.

(52) One first way to control the motor of the compressor 50 in this condition is illustrated in FIG. 16. FIGS. 14 and 15 shows two flow charts oriented to detect the overload mode in two different proposals of normal control. In this case, the overload control mode is configured to adjust the actuation frequency of the electric motor based on a piston dis-value de ((t)), or D.sub.MAX[K], with respect to the maximum reference displacement D.sub.REF. One observes that the function F illustrated in FIG. 14 (see second block A[k]=F(A[k−1],Ed[k]) may be a control P, PI or PID.

(53) In a second mode, as shown in FIG. 17, the overload control is configured to adjust the actuation frequency of the electric motor based on a velocity phase φ.sub.v of the motor of the compressor 50m, with respect a basis a velocity phase φv of the motor of the compressor 50m, with respect to a reference velocity φREF.

(54) A third way to adjust the actuation frequency of the compressor 50 is shown in FIG. 18. In this case, the overload control mode is configured to adjust the actuation frequency of the electric motor based on a value of the displacement phase φ.sub.d of the motor of the compressor, with respect to the reference displacement phase φ.sub.dREF

(55) Additionally, FIG. 19 shows an alternative way of adjusting the actuation frequency of said compressor 50. This is a way of controlling over load, configured to adjust the actuation frequency of the electric motor based on a minimum current phase value φ.sub.c.

(56) With regard to the above-described adjustment modes, they are given by the difference in phase between the piston displacement value (d.sub.e(t)) and an input voltage phase (V.sub.int.) preferably around −176 degrees (for the compressor defined by the parameters of Table 1). On the other hand, the adjustment of actuation frequency is given starting from the difference between the velocity phase value φv and an input voltage phase value Vint, preferably around −86 degrees (for the compressor defined by the parameters of Table 1).

(57) The present invention has, as an innovatory and differentiated characteristic over the prior art, a set of steps capable of adjusting the actuation frequency of the compressor 50 in an efficient and quite simplified manner for the overload control mode foreseen. Such a methodology takes into account the fact that said compressor comprises at least one electric motor, the latter being actuated by a frequency inverter. Said method comprises essentially the following steps:

(58) a) measuring and estimating, at every operation cycle T.sub.R of the resonant linear compressor 50, an actuation frequency F.sub.R, a maximum piston displacement d.sub.e(t) of the resonant linear compressor 50, and/or the piston displacement phase φd and/or the piston velocity phase φv and/or the current phase φc;

(59) b) comparing the maximum piston displacement d.sub.e((t) with a maximum reference displacement D.sub.REF, and calculating a displacement error Err;

(60) c) calculating an operation feed voltage value.sub.Am-pop of the electric motor, from an operation feed voltage value of previous cycle and of the displacement error Err obtained in the preceding step (s);

(61) d) comparing the operation feed voltage value A.sub.mpop of the electric motor calculated at the preceding step with a maximum feed voltage value Amax;

(62) e) if the operation feed voltage value A.sub.mpop calculated at step “c” is lower than or equal to the maximum feed voltage value A.sub.max, then deactivate an overload control mode of the electric motor and decrease the actuation frequency F.sub.R down to a mechanical resonance frequency; and returning to step a);

(63) f) if the operation feed voltage value A.sub.mpop calculated at step “c” is higher than the maximum feed voltage value A.sub.max, then activate the overload control mode and increase the actuation frequency F.sub.R up to an electromechanical resonance frequency.

(64) As to the first overload control mode, as illustrated in FIG. 16, one can state that it further comprises the following step:

(65) n) comparing the maximum piston displacement de(t) with a maximum piston displacement of a cycle d.sub.e(t−1) preceding the operation cycle T.sub.R;

(66) o) if the maximum piston displacement de(t) is higher than the piston displacement of the preceding cycle de(t), then comparing the actuation frequency F.sub.R with the actuation frequency of the preceding cycle F.sub.R((t−1);

(67) p) if the actuation frequency F.sub.R is higher than the actuation frequency of preceding cycle R.sub.R(t−1), then increasing the actuation frequency F.sub.R by a frequency delta value T.sub.f and returning to step a);

(68) q) if the actuation frequency F.sub.R is not higher than the actuation frequency of the preceding cycle F.sub.R(t−1), then decreasing the actuation frequency F.sub.R by a frequency delta value T.sub.f and returning to step a);

(69) r) if the maximum piston displacement d.sub.e(t) is not greater than the maximum piston displacement of preceding cycle d.sub.e(t−1), then comparing the actuation frequency F.sub.R with an actuation frequency of preceding cycle F.sub.R(t−1);

(70) s) if the actuation frequency F.sub.R is lower than that actuation frequency of preceding cycle F.sub.R(t−1), then increasing the actuation frequency F.sub.R by a frequency delta value T.sub.f and returning to step a);

(71) t) if the actuation frequency F.sub.R is not lower than the actuation frequency of preceding cycle F.sub.R(t−1), then decreasing the actuation frequency F.sub.R by a frequency delta value T.sub.f and returning to step a).

(72) It should be pointed out that steps “n” to “t” define an overload control mode for a maximum piston displacement value of the compressor 50.

(73) For the second overload control mode, as shown in FIG. 17, the following steps are foreseen:

(74) n) calculating a velocity phase φv of the piston of the compressor 50;

(75) o) comparing the velocity phase φv, calculated at the preceding step, with a reference velocity phase value φ.sub.VREF,

(76) p) if the velocity phase φv is higher than the reference velocity phase φ.sub.VREF, then increase the actuation frequency F.sub.R by a frequency delta value T.sub.f and returning to step a);

(77) q) if the velocity phase φv is not higher than the reference velocity phase φv.sub.VREF, then decrease the actuation frequency F.sub.R by a frequency delta value T.sub.f and returning to step a).

(78) for this second control mode, steps “n” to “q” define an overload control mode of the compressor 50 for an adjustment of reference velocity phase around −90 degrees (−86 for the compressor defined by the parameters of Table 1).

(79) A third way to adjust the actuation frequency, according to the teachings of the present invention, and as illustrated in FIG. 18, comprises the following steps:

(80) n) calculating a piston displacement phase φ.sub.d of the compressor 50;

(81) o) comparing the displacement phase φ.sub.d calculated at the preceding step with a reference displacement phase value φ.sub.DREF,

(82) p) if the displacement phase φd is higher than the reference displacement phase φ.sub.DREF, then increase the actuation frequency F.sub.R by a frequency delta value T.sub.f and returning to step a);

(83) q) if the displacement phase φd is not higher than the reference displacement phase φ.sub.DREF, then decrease the actuation frequency F.sub.R by a frequency delta value T.sub.f and returning to step a).

(84) The last steps “n” to “q” above define an overload control mode of the compressor 50 for an adjustment of reference displacement phase around −180 (−176 degrees for the compressor defined by the parameters of table 1).

(85) In turn, FIG. 19 shows a fourth way of adjusting the actuation frequency of the electric motor, consisting of the following steps:

(86) n) calculating a current phase φc of the compressor 50;

(87) o) comparing the current phase φc calculated at the preceding step with a current phase value φc−1 preceding the operation cycle TR;

(88) p) if the current phase φc is higher than the previous cycle current phase value φc−1, then comparing the actuation frequency F.sub.R with a previous cycle actuation frequency F.sub.R(t−1);

(89) q) if the actuation frequency F.sub.R is higher than the previous cycle actuation frequency F.sub.R(t−1), then increase the actuation frequency F.sub.R by a frequency delta value T.sub.f and returning to step a);

(90) r) if the actuation frequency F.sub.R is not higher than the previous cycle actuation frequency F.sub.R(−1), then decrease the actuation frequency F.sub.R by a frequency delta value T.sub.f and returning to step a);

(91) s) if the current phase value φc is not higher than the previous cycle current phase value φc−1, then comparing the actuation frequency F.sub.R with a previous cycle actuation frequency F.sub.R(t−1);

(92) t) if the actuation frequency F.sub.R is lower than the previous cycle actuation frequency F.sub.R(t−1), then increase the actuation frequency F.sub.R by a frequency delta value Tf and returning to step a);

(93) u) if the actuation frequency Fr is not lower than the previous cycle actuation frequency F.sub.R(t−1), then decrease the actuation frequency F.sub.R by a frequency delta value Tf and returning to step a);

(94) for steps “n” and “u” above, one defines an overload control mode of the compressor 50 for a minimum current shift.

(95) It should be pointed out that, as the piston displacement reaches the maximum reference value and reaches the resonance frequency again, the present system and method are configured to come out of the overload control.

(96) On the other hand, the present invention foresees a resonant linear compressor 50 provided with the presently designed actuation system and with the actuation method as defined in the claimed object.

(97) Finally, one can state that the actuation system and method for a resonant linear compressor 50 as described above achieve their objectives inasmuch as it is possible to increase the maximum power supplied to said compressor ion conditions of high load or overload for the same equipment design.

(98) Moreover, it should be pointed out that the present invention enables better preservation of the foods of the cooling equipment by increasing the maximum power supplied to said compressor. Further, it is possible, on the bases of the teachings of the invention, to reduce manufacture costs of the final product, as well as to increase the efficiency of the compressor 50 in its nominal operation condition, taking into account a better sizing of its linear actuator.

(99) A preferred example of embodiment having been described, one should understand that the scope of the present invention embraces other possible variations, being limited only by the contents of the accompanying claims, which include the possible equivalents.