CROSS REFERENCE TO RELATED APPLICATION

This application claims priority under 35 USC 119 to Brazilian Patent Application No. BR 102015016317-7 filed Jul. 7, 2015, and the entire disclosure of said Brazilian application is hereby incorporated by reference in its entirety into the present specification.

FIELD OF THE INVENTION

The present invention relates to a method and to a system for protecting a resonant linear compressor. More specifically, the present invention relates to a method and to a system configured so as to prevent the operation of a resonant linear compressor at a given drive frequency whose harmonic coincides with the structural resonance frequency of the compressor.

DESCRIPTION OF THE PRIOR ART

Alternating piston compressors generate pressure by compressing a gas inside a cylinder by means of the axial movement of a piston. In this regard, the gas existing in the outer part of the cylinder is in an area called low-pressure side (suction or evaporation pressure) and gets into the cylinder through a suction valve, where it is then compressed by the piston movement. After the gas has been compressed, it is expelled from the cylinder through a discharge valve to an area called high-pressure side (discharge or condensation pressure).

One of the types of alternating piston compressor is the resonant linear compressor. In this compressor model, the piston is actuated by a linear actuator, which comprises a support and magnets, being actuated by a coil and a spring, which associates the movable part (piston, support and magnets) to the fixed part (cylinder, stator, coil, head and frame). The movable parts and the spring form a resonant assembly of the compressor.

The resonant assembly actuated by the linear motor has the function of developing a linear alternating movement, causing the movement of the piston inside the cylinder to exert a compression action of the gas admitted through the suction valve as far as the point where it is discharged through the discharge valve.

For this reason, amplitude of operation of the resonant linear compressor is regulated by the balance of the power generated by the motor and the power consumed by the mechanism in the compression, besides the losses generated in this process. Thus, in order to achieve maximum thermodynamic efficiency, resulting in maximum cooling capacity, the piston displacement should draw near to the stroke end (as close to the head as possible), so as to reduce the volume of dead gas (unused gas) in the compression process.

Thus, in order to make the compression process feasible with maximum efficiency, it is necessary to have precision in the analysis and knowledge of the piston stroke, preventing the risk of impact of the piston against the stroke end, which would generate acoustic noise, loss of efficiency and even a possible break of the resonant linear compressor.

So, the greater the error in detecting the piston stroke the greater the safety coefficient necessary between the maximum piston displacement and the stroke end, increasing losses of output in the product.

On the other hand, the system has lesser need for cooling, and so it is necessary to reduce the cooling capacity of the resonant linear compressor. It is possible to reduce the power stroke of the piston, thus diminishing the power supplied to the system, promoting a variable cooling capacity of the compressor, which may be controlled by controlling the piston stroke.

Besides, another important characteristic of resonant linear compressors is the drive frequency. The system in which such compressor are used are designed to operate at a specific resonance frequency of the mass/spring system, since at this point the reactive forces of the system are annulled and, as a result, the system reaches maximum efficiency. Such drive frequency is derived from the actuation of the spring of the resonant linear compressor and from the amplitude A of the Aa feed voltage on the piston.

By “mass/spring” one understands that mass (m) the sum of the mass of the movable part (piston, support and magnet) and the equivalent spring (K_{T}) is the sum of the resonant spring of the system (K_{ML}) plus the gas-compression force which, since it is dependent upon the evaporation and condensation pressures of the cooling system, as well as of the gas used for compression, may be modeled to one more spring constant (K_{G}).

Such theories can be found in papers of the *IEEE*, as for example, “*A Novel Strategy of Efficiency Control for a Linear Compressor System Driven by a PWM Inverter*” (by authors T. Chun, J. Ahn, H. Lee, H. Kim and E. Nho), as well as “*Method of Estimating the Stroke of LPMSM Driven by PWM Inverter in a Linear Compressor*” (by authors T. Chun, J. Ahn, Q. Tran, H. Lee and H. Kim), “*Analysis and control for linear compressor system driven by PWM inverter*” (by authors T. Chun, J. Ahn, J. Yoo and C. Lee) and “*Analysis for sensorless linear compressor using linear pulse motor*” (by authors M. Sanada, S. Morimoto and Y. Takeda).

In this regard, the paper “*A Resonant Frequency Tracking Technique for Linear Vapor Compressors*” (by authors Z. Lin, J. Wang and D. Howe) presents another theory that such mass/spring systems can calculate a resonance frequency (f_{r}) by the equations (1) and (2) below:

*K*_{T}*=K*_{ML}*+K*_{G} (1)

$\begin{array}{cc}{f}_{R}=\frac{1}{2.\pi}\sqrt{\frac{{K}_{T}}{m}}& \left(2\right)\end{array}$

Since the spring gas portion is unknown (K_{G}), non-linear and variable throughout the operation of the resonant linear compressor, it is not possible to calculate the resonance frequency with the precision necessary to optimize the efficiency of this type of compressor. This paper also presents a theory of adjusting resonance frequency, where one applies a variation of drive frequency as far as the maximum power point, for a constant current, thus presenting a simple and easy-to-implement method, which, however, needs to disturb the system periodically to detect the resonance frequency.

Further, as can be seen in the already cited papers and additionally in document WO0079671, when the system operates at the resonance frequency, the motor current is in quadrature with the displacement, that it, the motor current is in phase with the counter-electromotive force (CEMF, or back-EMF) of the motor (considering that the CEMF is proportional and derived from the displacement). This method is more precise to optimize the efficiency of the compressor, but it needs constant detection of the current phase and of the displacement phase, thus needing position or velocity sensing cars.

If the structural resonance frequencies are excited, this originates disturbances in the functioning of the resonant linear compressor, which may vary from the increase in acoustic noise to the break thereof. Therefore, control methods are necessary so that such (structural resonance) frequencies will not be excited or, alternatively, methods that prevent the resonant linear compressor from operating at such frequencies. One of the viable approaches is the mechanical modification in the compressor construction, so that the structural resonance frequencies will be outside the area of the harmonic of the main resonance frequency of the system.

However, due to the variability of the productive process and of the variation in the main resonance frequency (due to variation of the charge), it may not be possible to prevent harmonics of the drive frequency from exciting structural resonances.

Thus, another approach would be to prevent the drive of the system at frequencies that have harmonics that excite the structural resonance frequencies. This solution may lead to a minor drop in efficiency of the system, due to the fact that the compressor is not actuated exactly at the resonance frequency (when a harmonic of the later coincides with a structural resonance), but, on the other hand, this guarantees the reliability and durability of the compressor.

Solutions to this problem appear only on rotary motors, as shown, for instance, in document U.S. Pat. No. 5,428,965, which describes a control system for variable-speed motors, which prevents drive of the motor at certain velocities to prevent excessive noise or vibrations, or document EP 2,023,480, which describes the control of rotary motors that modifies the current phase to prevent drive at these frequencies, reducing the noise and vibrations of the motor.

These techniques, however, are not easy to apply for linear motors. On rotary motors there is a control over the frequency of operation of the compressor, that is, one can vary the operation frequency without concerns relating to losses of the system.

Thus, rotary motors have an effect that is totally different from that of linear motors. As already explained, electric motors that have magnets produce a force that is contrary to motion force of the motor, called counter-electromotive force (CEMF). This CEMF ends up limiting the voltage (and, as a result, the current that is applied to the motor. So, modifying the phase of the current applied on rotary motors with respect to the CEMF makes the application of a higher current with respect to the phase with the CEMF (called also field suppression on rotary machines) impossible. Since the frequencies of these compressor is determined only by the motor, a rotary compressor can modify the operation frequency upon modifying the frequency of its inverter, without any concern with loss of efficiency, since its energy is constant, always determined by the value of the kinetic energy.

This effect, however, is different for resonant linear machines, the later operating at the main resonance frequency of the system, this being the function of the product design, which may undergo minor variations due to the gas compression effect.

Factors like the temperature in the environment in which the compressor is arranged may also interfere with the main resonance frequency of the system. For instance, in cold environments the main resonance frequency of the resonant compressor is at 110 Hertz. On the other hand, in a warmer environment, as the discharge pressure of the compressor increases, the main resonance frequency reaches 130 Hertz.

In other words, there is no control over the operation frequency of the compressor, so that this frequency may vary in a short period of time (due to weather variations).

During the movement of resonant motors, there is a constant change of kinetic energy and potential energy, the resonance frequency being the point at which the kinetic energy and the potential energy have the same amplitude. At this frequency, when the piston is at its maximum speed, the kinetic energy represents the whole energy of the system, whereas at the uppermost or lowermost points (top or bottom dead center), the potential energy represents the whole energy of the system and the total energy of the system is always constant, oscillating between kinetic and potential energy.

Upon modifying the frequency, that is, upon getting out of the resonance, the potential energy or the kinetic energy will prevail in the system, and the additional energy to keep the balance (and the functioning of the system) shall be produced by an external system, which in this case is the motor. In this way, if the operation frequency on a resonant linear compressor is different from the main resonance frequency, the motor of this compressor will view a relative charge that is additional to the system, which does not generates work, but consumes energy (in this case, accelerating and decelerating the piston, which at the resonance frequency is carried out automatically by the spring in the exact extent to annul any reactive charge).

Since linear compressor should always operate at the resonance frequency, factors like variations in the charge or temperature may modify the operation frequency, and this frequency should be accompanied by the inverter of the motor, for better drive efficiency.

Thus, modification of frequency on linear machines may not be considered obvious with respect to modification on rotary machines, since on linear compressors modification in the frequency (operation of the compressor out of resonance) will generate reactive loads which must be absorbed by the compressor motor. On rotary compressors, as already mentioned, the variation in the frequency does not entail great losses for the system.

Thus, there is no description, in the prior art, of a method or a simple and useful system that prevents the operation of a resonant linear compressor at drive frequencies whose harmonics coincide with the structural resonance frequency of the system.

BRIEF DESCRIPTION OF THE INVENTION

This application describes a method for protecting a resonant linear compressor, such a compressor comprising structural resonance frequency and a motor that is fed by a feed voltage that exhibits an amplitude and a drive frequency, both controlled according to the equation A·sin(wt).

The protection method is configured so as to comprise a step of preventing feed to the motor at drive frequencies that have at least one harmonic coinciding with the structural resonance frequency of the resonance linear compressor.

The present invention further relates to a system for protecting a resonant linear compressor, which comprises an electronic control and is configured so as to prevent feed to the motor at drive frequencies that have at least one harmonic coinciding with the structural resonance of the resonant linear compressor.

BRIEF DESCRIPTION OF THE DRAWINGS

The present invention will be described in greater detail with reference to an embodiment represented in the drawings. The figures show:

FIG. 1—is a cross-sectional view of a resonant linear compressor;

FIG. 2—is a mechanic model of the resonant linear compressor;

FIG. 3—is an electric model of the resonant linear compressor;

FIG. 4—is a response diagram at frequency of the function of displacement transfer of the mechanical system;

FIG. 5—is a response diagram at frequency of the velocity of the mechanical system;

FIG. 6—represents a graph of the drive frequency (Hertz) of the resonant linear compressor as a function of its vibration;

FIG. 7—represents a graph of the drive frequency (Hertz) of the resonant linear compressor as a function of its vibration;

FIG. 8—represents a time graph (seconds) as a function of the drive frequency (Hertz) of a resonant linear compressor;

FIG. 9—is a time graph (seconds) as a function of the current (amperes) indicating the ideal condition of operation of a resonant linear compressor;

FIG. 10—is a graph representing the control of the drive frequency of the resonant linear compressor upon delaying the current phase;

FIG. 11—is a graph representing the control of the drive frequency of the resonant linear compressor upon advancing the current phase;

FIG. 12—is a representation of the drive frequency of the resonant linear compressor as a function of the phase between the electric current and the piston displacement velocity;

FIG. 13—represents a flowchart describing the “phase jump” according to the method proposed in the present invention;

FIG. 14—is a representation of the drive period of the resonant linear compressor as a function of the phase between the piston velocity and the electric current;

FIG. 15—represents a flowchart describing the “phase jump” according to the method proposed in the present invention, considering the drive period of the resonant linear compressor;

FIG. 16—is block representation of the system for protecting a resonant linear compressor as proposed in the present invention.

DETAILED DESCRIPTION OF THE FIGURES

FIG. 1 illustrates the embodiment of the resonant linear compressor **14**, in which the system and the method proposed in the present invention are applied. For a better understanding of the figures, the resonant linear compressor **14** will be described only as compressor **14**, in a few situations.

Said compressor **14** comprises a piston **1**, a cylinder **2**, a suction valve **3***a *and a discharge valve **3***b*, besides having also a linear actuator comprising a support **4** and magnets **5**, the latter being actuated by one or more coils **6**.

The resonant linear compressor **14** further has one or more springs **7***a *and **7***b*, which connect a movable part of the compressor **14**, comprising the piston **1**, the support **4** and the magnets **5**, a fixed part of the compressor **14**, comprising the cylinder **2**, a head **3**, at least one stator **12**, to which the coils **6** are fixed, and a structure **13** for fixation of all the elements necessary for the correct operation of the compressor **14**.

During the operation of the compressor **14**, the gas gets into the cylinder **2** through the suction valve **3***a *and is compressed by a linear movement of the piston **1**, being later expelled from the system by the discharge valve **3***b*. The movement of the piston **1** in the cylinder **2** is made by actuation of the coils **6** of the stator **12** on the magnets **5** associated to the support **4**, besides the opposite movement made by actuation of the springs **7***a *and **7***b *on the same support **4**.

In this regard, FIG. 2 presents a mechanical model of the compressor **14** (mass/spring mechanical system) of FIG. 1, wherein equation (3) can be obtained (3).

$\begin{array}{cc}m\xb7\frac{{\delta}^{2}d\left(t\right)}{\delta \phantom{\rule{0.3em}{0.3ex}}{t}^{2}}={F}_{\mathrm{MT}}\left(i\left(t\right)\right)-{F}_{\mathrm{ML}}\left(d\left(t\right)\right)-{F}_{\mathrm{AM}}\left(v\left(t\right)\right)-{F}_{G}\left(d\left(t\right)\right)& \left(3\right)\end{array}$

In equation (3), the motor force in Newton is defined by F_{MT}(i(t))=K_{MT}·i(t), whereas the spring force, also in Newton, defined by F_{ML}(d(t))=K_{ML}·d(t). The dumping tons is modeled or F_{AM}(ν(t))=K_{AM}·ν(t) and similarly the gas-pressure force within the cylinder, again in Newton, is defined by F_{G}(d(t)). In these equations, K_{MT }is the modeling of a spring constant of the motor (motor constant), whereas K_{ML }is the é the spring constant and K_{AM }represents the modeling of the damping constant.

The mass of the movable part of the system is defined by m, the piston velocity being defined by ν(t), the piston displacement by d(t) and the current in the motor by i(t).

FIG. 3 shows an electric modeling (RL electric circuit in series with a strong voltage) of the compressor **14** of FIG. 1, in which one can obtain the equation (4).

*V*_{ENT}(*t*)=*V*_{R}(*i*(*t*))+*V*_{L}(*i*(*t*))+*V*_{MT}(ν(*t*)) (4)

In this equation (4), the voltage of the resistance in Volts is modeled by V_{R}(i(t))=R·i(t), wherein R is the electric resistance of the motor. On the other hand, the inductor voltage, also in volts, is modeled by

${V}_{L}\left(i\left(t\right)\right)=L\xb7\frac{\mathrm{di}\left(t\right)}{\mathrm{dt}},$

wherein L represents the motor inductance.

The voltage induced in the motor (CEMF) in Volts is represented by V_{MT}(ν(t))=K_{MT}·ν(t), whereas the feed voltage, also in Volts, is represented by V_{ENT}(t).

The gas-pressure force F_{G}(d(t)) is not constant, the latter being variable as a function of the changes in suction pressure and discharge pressure and, as a result, with piston displacement.

The other forces in the mechanical equation (mass/spring modeling), as well as all the voltages of the electric equation (RL circuit), are linear functions. In order for us to achieve a complete model of the system, it is possible to replace the pressure force by the modeled effects which it causes in the system, said effects being the consumption of power and the variation in the resonance frequency.

The consumption of power may be modeled by an equivalent (variable) damping, whereas the variation in the resonance frequency is modeled by an equivalent spring (also variable).

Thus, the equation (3) may be re-written according to the equation (5) or (6) bellow.

$\begin{array}{cc}m\xb7\frac{{\delta}^{2}d\left(t\right)}{\delta \phantom{\rule{0.3em}{0.3ex}}{t}^{2}}={K}_{\mathrm{MT}}\xb7i\left(t\right)-\left({K}_{\mathrm{ML}}-{K}_{\mathrm{MLEq}}\right)\xb7d\left(t\right)-\left({K}_{\mathrm{AM}}+{K}_{\mathrm{AMEq}}\right)-v\left(t\right)& \left(5\right)\\ \phantom{\rule{4.2em}{4.2ex}}m\xb7\frac{{\delta}^{2}d\left(t\right)}{\delta \phantom{\rule{0.3em}{0.3ex}}{t}^{2}}={K}_{\mathrm{MT}}\xb7i\left(t\right)-{K}_{\mathrm{MLT}}\xb7d\left(t\right)-{K}_{\mathrm{AMT}}\xb7v\left(t\right)& \left(6\right)\end{array}$

In these equations (5) and (6), K_{MLEq }determines the modeled coefficient of the equivalent spring, whereas K_{AMEq }represents the equivalent damping equivalent. The total spring coefficient, K_{MLT}, may be calculated as K_{MLT}=K_{ML}+K_{MLEq}.

In the same way, the total damping coefficient may be calculated as K_{AMT}=K_{AM}+K_{AMEq}. Thus, upon applying the Laplace transform to equations (4) and (6) it is possible to obtain the equation (7), which represents the electric equation in the frequency domain, besides the mechanical equations (8) and (9), which represent the transfer function between the displacement and the velocity relating to the current, as shown below:

$\begin{array}{cc}I\left(s\right)=\frac{{V}_{\mathrm{ENT}}\left(s\right)-{K}_{\mathrm{MT}}.V\left(s\right)}{L.S+R}& \left(7\right)\\ \frac{D\left(s\right)}{I\left(s\right)}=\frac{{K}_{\mathrm{MT}}}{m.{s}^{2}+{K}_{\mathrm{AMT}}.s+{K}_{\mathrm{MLT}}}& \left(8\right)\\ \frac{V\left(s\right)}{I\left(s\right)}=\frac{{K}_{\mathrm{MT}}.s}{m.{s}^{2}+{K}_{\mathrm{AMT}}.s+{K}_{\mathrm{MLT}}}& \left(9\right)\end{array}$

Thus, the mechanical resonance frequency is given by the module of the pair of complex poles of the equation characteristic of the mechanical system, this being the frequency at which the system exhibit better relation between current and displacement (or velocity), that is higher efficiency.

FIGS. 4 and 5 show reply diagrams at frequency (Bode diagrams) of the transfer function of the displacement of the mechanical system (FIG. 4) and of the velocity of the mechanical system (FIG. 5). In these figures, one observes that at the mechanical resonance frequency the system gain is maximum (maximum magnitude). Further, the displacement is offset 90 degrees with respect to the current (displacement and current are in quadrature) and the velocity is in phase with respect to the current (phase between velocity and current is of 0 degree).

Thus, the variations in load may be represented by variations in the total spring coefficient and in the total damping coefficient, these factors will affect the resonance frequency and the gains of the system.

The structural resonances may be represented as a mass/spring system, as in FIG. 2 and conforming to the equation (3), but without undergoing influence of the load and depending only on the dimension characteristics of the compressor **14**. In other words, the structural resonance is constant for the same compressor **14** (even considering variations in temperature), but it varies between different compressors, that is, the structural resonance is never identical.

Because of this, the structural resonance exhibit low dampening and a high spring constant, so that their (structural) resonance frequency is considerably higher than the main resonance frequency of the system, being possible located on harmonics of the main resonance frequency of the system (drive frequency).

Thus, and just as mentioned before, the operation of the linear compressor **14** at the structural resonance frequencies may entail damage to the compressor **14**, so that it is advisable that the functioning of the compressor **14** at such frequency should be prevented.

In this regard, the present invention discloses a method and a system for protecting a resonant linear compressor **14** which have the objective of preventing the operation of the compressor **14** at the structural resonance frequency of the system. In other words, the present invention relates to a method and to a system for protecting a resonant linear compressor **14** which prevent harmonics of the drive frequency from coinciding with the structural resonance of the system.

Such a resonant linear compressor **14** comprises structural resonance frequencies w_{E }and a motor, the latter being fed by a feed voltage Va provided with amplitude A and a drive frequency w_{A}, both controlled according to the equation A·sin(wt).

FIGS. 8 and 7 show a graph of the drive frequency of the linear compressor **14** as a function of its variation. One observes in FIG. 6 that the third harmonic of the drive frequency w_{A }is above the structural resonance of the system.

The situation that one wishes to prevent in order to protect the linear compressor **14** and the system which it integrates is shown in FIG. 7. In this case, one observes that the third harmonic of the drive frequency w_{A }is equal (coincides with) to the structural resonance of the system, which entails excess vibration to the resonant linear compressor **14**.

In order to prevent operation of the resonant linear compressor **14** at harmonics of the drive frequency w_{A }from coinciding with the structural resonance frequency w_{E }of the system, one starts from the presupposition that the latter is known. For this purpose, for instance, one can detect the counter-electromotive force of the linear actuator or still use a sensor for sensing position or velocity of the piston of the resonant linear compressor **14**.

In the method and in the system for protecting a resonant linear compressor **14**, as proposed in the present invention, one considers a resonant linear compressor **14** in which one knows that the structural resonance frequency w_{E }coincides with the third harmonic of the drive frequency, as shown in FIG. 7.

FIG. 8 shows a time graph (seconds) as a function of the drive frequency w_{A}, at Hertz, of the resonant linear compressor **14**. One observes that in this situation the drive frequency of the compressor **14** drops as a function of the time. As already mentioned, such a situation may occur due to the drop in temperature of the environment in which the compressor **14** is arranged.

Thus, during the variation in drive frequency w_{A }of the compressor, it may happen that a harmonic of the drive frequency w_{A }coincides with the structural resonance frequency w_{E}, a situation which, as already mentioned, one wishes to prevent.

The structural resonance frequency w_{E }of the compressor **14** is indicated from the dashed line of the operation frequency w_{A}. One observes that such a frequency coincides with the third harmonic of the drive frequency 3*w_{A}. Thus, it is desirable to prevent the drive of the compressor at the drive frequency w_{A }coinciding with the structural resonance frequency w_{E}.

For this purpose, the method for protecting a resonant linear compressor **14** as proposed in the present invention alters the drive frequency w_{A }by varying the phase between the electric current i(t) of the compressor **14** and the velocity of piston displacement. In this way, the efficiency of the compressor is slightly impaired. On the other hand, noises and excess disturbances are prevented on it.

Knowing the structural resonance frequency w_{E }of the system, an electronic control of the linear compressor **14**, upon detecting a point higher than 10 of the structural resonance frequency w_{E}, will advance the phase between the electric current i(t) of the compressor **14** and the velocity of piston displacement.

Upon reaching the point at which the phase may not be offset any longer (minimum offset value **12**), the later should be delayed and will later return to phase 0°, thus causing a “frequency jump”. This frequency jump will jump over the structural resonance frequency w_{E }of the system, thus preventing the noises and vibrations that may damage the linear compressor **14**.

In a similar way, this jump in the structural resonance frequency C_{fase }is carried out if the linear compressor **14** is arranged in an environment in which the room temperature is rising. In this situation, the electronic control, upon detecting a lower point **11** of the structural resonance frequency w_{E }will delay the phase between the current and the displacement until the maximum offset value **15** and then will reestablish it and later return to the phase 0°, thus causing said “jump” in the structural resonance frequency w_{E}.

FIGS. 9, 10 and 11 represent a graph of the time (seconds) as a function of the current (amperes) of the linear compressor **14**. FIG. 9 represents the ideal functioning condition of said compressor **14** (compressor **14** operating perfectly at the resonance, that is, actuating symmetrically in the two directions of piston displacement), this situation being represented in FIG. 9 and indicating the operation of the compressor **14** out of the structural resonance frequency w_{E}.

The delay in the offset of the current is indicated in the graph of FIG. 10, in which one observes that the end of the current gets close to the upper dead center (UDC) and to the lower dead center (LDC) of the piston displacement. On the other hand, the operation frequency of the compressor **14** is lower if compared with the operation frequency indicated in FIG. 9.

The graph shown in FIG. 11 represents the current advanced in phase if compared with the graph in FIG. 10. In this situation, the start of the current gets close to the PMS and PMI and the operation frequency of the compressor **14** is higher is compared with the frequency indicated in FIG. 10.

It is valid to mention that, although this preferred embodiment of the present invention describes this jump in the structural resonance frequency C_{fase }for the third harmonic of the drive frequency, in another linear compressor, this “jump” in the frequency might occur, for example, in the fourth harmonic.

Additionally, FIG. 12 a representation of the frequency of the linear compressor **14** as a function of the phase between the electric current i(t) and the piston velocity. As in the graph shown in FIG. 8, but now shown in the so-called hysteresis signal, FIG. 12 shows the phase control for preventing drive of the compressor **14** at the structural resonance frequency w_{E }of the system.

In this graph and more precisely at the abscissa axis, one represents a lower limit and an upper limit for the structural resonance frequency w_{E}, called F_{rLI }and F_{rLS}, respectively. Thus, in the regions in which the drive frequency w_{A }of the compressor **14** is F_{rLI}<w_{A}<F_{rLS}, region is configured in which one wishes to prevent drive of the compressor **14**, that is, the region in which said “frequency jump” will take place.

On the other hand, the ordinate axis refers to the phase between the current and the velocity and the graph shown in FIG. 12, represents a first lower limit of the phase F_{sLI1}, a second lower limit of the phase F_{sLI2}, a first upper limit of the phase F_{sLS1 }and a second upper limit of the phase F_{sLS2}.

FIG. 13 represents a flowchart describing the “phase jump” shown in the graph of FIG. 11. One observes that at the start of a new cycle of piston displacement **1**, the decision step **20** verifies whether (w_{A}<F_{rLS}) and (w_{A}>w_{E}), which indicates the region between w_{E }and F_{rLS }(FIG. 12). If so, the decision step **21** verifies whether F_{s}>Fs_{LI2 }and, if so, the phase between the current and the velocity will be advances (operation step **22**), assuming the velocity as a reference.

If not, the phase F_{s }will be reestablished, assuming the value of the F_{sLs1}, as shown in FIG. 12.

If the step **20** give a negative result, the condition step **23** will verify whether (w_{A}>F_{rLI}) and (w_{A}<w_{E}), which would represent the region between F_{rLI }and w_{E }(FIG. 12). In this case, the condition step verifies whether F_{s}<F_{sLS2}, if so, the phase of the current with respect to the velocity will be delayed, according to the operation step **25**. It not, the current phase will be reestablished, assuming the value of F_{sLI1}, as shown in FIG. 12.

Thus, the phase values of the second lower limit F_{sLI2 }and of the second upper limit F_{sLS2 }represent the minimum and maximum offset values, respectively, so that, for values lower than F_{sLI2 }(second lower limit) such offsetting will be reestablished (assuming the value of F_{sLs1}), and, in a similar way, for values hither than F_{sLS2 }(second upper limit) the offsetting is reestablished, assuming the value of F_{sLI1 }(first lower limit).

The minimum and maximum offsetting value F_{sLI2}, F_{sLS2 }are related to the moment when the drive current of the compressor is zero, moments when the points PMS and PMI (FIG. 9) are detected and when, as a result, the counter-electromotive force generated by the motor is also null.

Following the description of the flowchart shown in FIG. 13, if the conditions steps **20** and **23** assume negative conditions, which would represent operation of the compressor **14** out of the limits of the structural resonance frequency w_{E }(normal operation of the compressor), in this case the condition step **26** verifies whether the phase F_{s }will be delayed, according to step **27**. If not, the condition step **28** verifies whether F_{s}>0 and, if positive, the phase F_{s }is advanced, if not, the cycle reaches its end.

Specifically, the “phase jump” is shown at steps **20** to **25**, which take as a basis the verification of the drive frequency w_{A}. Steps **26** and **28** refer to the normal operation of the compressor (w_{A}<F_{rLI }or w_{A}>F_{rLS}), and in this condition the phase F_{S }(phase between the current and the displacement velocity) should be kept 0°.

For this reason, the condition step **26** delays the phase F_{s }if F_{s}<0 and the condition **28** advances the phase F_{s }if F_{s}>0, that is, such steps cause the offsetting to be equal to 0°, equivalent to the condition of normal operation of the compressor, thus guaranteeing the perfect operation tuning thereof.

Thus, the operation of the compressor **14** at the structural resonance frequency w_{E }(F_{rLI}<w_{E}<F_{rLS}) will be prevented. Further, a new cycle will be started from the step **20** whenever the piston **1** reaches its upper dead center PMS or lower dead center PMNI (FIGS. 9, 10, and 11).

In a numerical example of said “phase jump” shown in FIG. 12, supposing that the phase Fs is at 0° and the lower limit F_{rLI }of the structural resonance frequency is detected (due to the rise in temperature at which the compressor is arranged), the phase Fs will be delayed to 20° (F_{sLs2}) and then reestablished to −15° (F_{sLI1}), at the moment when the upper limit of the structural resonance frequency F_{rLS }is detected, the phase will again be delayed to 0°. Obviously, such values are only preferred features of the present invention and should not be considered compulsory.

In a similar way, and considering now a drop in temperature of the environment where the compressor id arranged, upon detecting the upper limit F_{rLS }of the structural resonance frequency, the phase Fs will assume the value −20° (F_{sLI2}) and then reestablished to 15° (F_{sLs1}).

The reason why the graph in FIG. 12 discloses two levels of “phase jump”—a first level being composed by the points F_{sLs2 }and F_{sLI1 }and a second level formed by the points F_{sLs1 }and F_{sLI2}—would be to prevent instability at the moment of the “jump”, so that in the cases where only one level is used the occurrence of minor noises may entail indecision about which is the correct value of the phase which should be established.

These two level of phase jump are called levels of hysteresis and, in this preferred example, there is a hysteresis of 5°, since the first upper limit F_{sLs1 }and the second upper limit F_{sLs2 }assume preferable values of 15° and 20°, respectively.

It is important to mention that if the “phase jump” does not comprise the levels of hysteresis shown in FIG. 8 of the present application, in this case the maximum and minimum values of offsetting **15**, **10** will be preferably 20° and −20°, respectively.

One can then establish an analogy between the graphs of FIGS. 8 and 12, in which the upper point **10** is equivalent to the upper limit F_{rLS}, the lower point **11** is equivalent to the lower limit F_{rLI}, the maximum offsetting **15** is equivalent to F_{sLs2 }and the minimum offsetting value **12** is equivalent to F_{sLI2}.

In an additional embodiment of the present invention, the operation of the resonant linear compressor **14** may be interrupted, if it is found that the drive frequency w_{A }comprises values higher than F_{rLI},**11** and lower than F_{rLS},**10**, that is, the lower limit and upper limit (respectively of the structural resonance frequency w_{E}.

Further, the graph shown in FIG. 14 and the flowchart of FIG. 15 are analogous to those represented in FIGS. 12 and 13, respectively. More specifically, FIG. 14 represents a graph of the period with respect to the phase between the current and the velocity.

In this graph, instead of the structural resonance frequency w_{E}, a structural resonance period t_{E }is represented, delimited by a lower limit T_{LI }and an upper limit T_{LS}. On the other hand, the flowchart of FIG. 15 represents the control of the phase by the period from a drive period t_{A}. The steps exhibited in this flowchart are equivalent to those shown in FIG. 13, but it takes into consideration the period, not the drive frequency w_{A }of the compressor **14**.

The present invention further relates to a system for protecting a resonant linear compressor **14** capable of carrying out the method proposed in the present invention. In other words, said system is configured so as to prevent feed of the linear compressor at drive frequency w_{A }whose harmonics coincide with the structural resonance frequency w_{E }of the compressor **14**.

As can be observed from FIG. 16, said protection system is provided with an electronic control **30**, the latter comprising at least one rectifier **31**, one control unit **32** and one converter **33**. The proposed system, by means of its electronic control **30**, is capable of measuring the electric current i(t) of the motor, calculating the phase thereof, as well as a period of an operation cycle. Further, the system is configured so as to measure or estimate the displacement or the velocity of the piston, as well as calculating the phase thereof, and is further capable of measuring the counter-electromotive force of the linear compressor **14**.

Additionally, the protection system proposed in the present invention is configured so as to advance or delay the phase between the electric current i(t) of the compressor **14** and the piston displacement velocity, if at least one harmonic of the drive frequency w_{A }coincides with the structural resonance frequency w_{E }of the resonant linear compressor **14**, as can be observed in FIGS. 8 to 12 of the present invention.

Said protection system is further capable of reestablishing the phase between the electric current i(t) of the compressor and the piston displacement velocity, if the latter assumes values lower than the minimum offsetting value F_{sLI2},**12** or values higher than the maximum offsetting value F_{sLS2},**15**, as shown in FIG. 12.

The proposed system is further capable of reestablishing the phase between the electric current i(t) of the compressor **14** and the piston displacement velocity, from a second upper limit F_{sLS2 }to a first lower limit F_{sLI1 }and from a second lower limit F_{sLI2 }to an first upper limit F_{sLS1}.

In an alternative configuration of the present invention, the protection system is further configured so as to interrupt the electric drive of the resonant linear compressor **14**, if the electronic control **30** verifies that the drive frequency w_{A }assumes values higher than a lower limit value F_{rLI},**11** and lower than an upper limit value F_{rLS},**10** of the structural resonance frequency w_{E}.

In other words, the proposed system can, instead of making the so-called “frequency jump”, interrupt the operation of the linear compressor **14**, if it is verified that the latter is at operation at a drive frequency w_{A }that coincides with the structural resonance frequency w_{E }of the compressor **14**.

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.